CN113646518A - Separately determining ignition density and pumping density during ignition density transitions for lean-burn internal combustion engines - Google Patents

Abstract

A skip fire engine controller and control method are described in which, during a transition from a first firing density to a second firing density, the firing density and the pumping density are individually set so as to balance conflicting requirements of: (a) torque control, (b) noise, vibration and harshness (NVH), (c) air flow through the engine, and (d) air fuel ratio.

Description

Separately determining ignition density and pumping density during ignition density transitions for lean-burn internal combustion engines

Cross Reference to Related Applications

This application claims priority to U.S. application serial No. 16/576,972 filed on 20/9/2019. This application is also a partially-filed and priority-filed us application No. 16/373,364(TULA P076) on day 4/2 in 2019. The two applications listed above are incorporated by reference herein in their entirety for all purposes.

Technical Field

The present invention relates to a skip fire engine control system for a lean burn internal combustion engine, and more particularly to skip fire engine control in which an ignition mode and a pumping mode are implemented separately during a transition from a first ignition density to a second ignition density in order to manage and balance conflicting demands of: (a) satisfying a torque request, (b) preventing excessive noise, vibration, and harshness (NVH), (c) maintaining a proper air-fuel ratio, and (d) controlling air flow through the engine.

Background

Most vehicles in operation today are powered by Internal Combustion (IC) engines. Internal combustion engines typically have a plurality of working chambers (i.e., cylinders) in which combustion occurs. The power generated by the engine depends on a combination of both: (a) the number of cylinders, and (b) the amount of fuel and air delivered to each cylinder. During routine driving, the engine of the vehicle is typically operated over a wide range of torque demands and operating speeds to meet different driving conditions.

There are two common types of IC engines; spark Ignition (SI) engines and compression ignition engines. Both engine types typically use a cylinder as a working chamber, in which a piston reciprocates, forming a variable sized enclosed volume depending on the piston position. Air is directed from the intake manifold into one or more cylinders through one or more intake valves by pushing a piston to expand a closed volume. The guided air is then compressed by the piston movement in order to contract the enclosed volume. Combustion occurs within the cylinder's containment volume at or near its minimum size. The expanding combustion gases push the piston outward, expanding the enclosed volume and performing useful work. The piston in turn pushes exhaust gas out of the enclosed volume through one or more exhaust valves into an exhaust manifold.

For SI engines, combustion is initiated by a spark. That is, an air-fuel mixture is contained within the cylinder(s) of the engine and a spark (typically from a spark plug) is then used to ignite the mixture.

Compression engines, on the other hand, rely on the pressure and temperature of the air-fuel mixture to initiate combustion rather than a spark. The air-fuel mixture is contained within the cylinder, and the mixture is raised in temperature by mechanical compression to cause combustion, resulting in auto-ignition of the fuel.

The air-fuel ratio in the cylinder is an important measure for both SI engines and compression ignition engines. This ratio is referred to as "stoichiometric" if sufficient air is provided to completely combust all of the fuel without any remaining oxygen.

A sub-stoichiometric ratio is considered "rich," meaning that the ratio defines more fuel than the amount of air that can be combusted. The rich mixture can produce more power and burn at a lower temperature, but at the expense of efficiency.

On the other hand, a ratio above stoichiometry is considered "lean," meaning that the ratio defines an air-fuel mixture having more oxygen than the fuel can combust. Lean air-fuel ratios do not effectively use a common three-way catalyst in an exhaust aftertreatment system because excess oxygen is typically present in the exhaust.

Spark-ignition engines generally operate at a stoichiometric fuel/air ratio and their output torque is controlled by controlling the Mass Air Charge (MAC) in the cylinders. Mass air charges generally use throttle control to reduce intake Manifold Absolute Pressure (MAP). Spark ignition engines may also use a super or turbocharger to boost intake manifold pressure above atmospheric pressure.

Compression ignition engines typically control engine output torque by controlling the amount of fuel injected (thus changing the air/fuel ratio) rather than the air flow through the engine. Engine output torque is reduced by adding less fuel to the air entering the cylinders (i.e., running the engine leaner). Compression ignition engines generally operate at lean air/fuel ratios. For example, a diesel engine (which is the most common type of compression ignition engine) may typically be operated at an air/fuel ratio in the range 16 to 55, as compared to a stoichiometric air/fuel ratio of approximately 14.6. Some diesel engines, which are typically older, do not generally use a throttle valve, but instead use a turbocharger to control the flow of air into the engine. Compression-ignition engines may be further classified based on their fuel and the manner in which the fuel is mixed with air within the cylinder. Several common types of compression ignition engines include: stratified charge compression ignition engines (e.g., the most conventional diesel engines and abbreviated as SCCI), Premixed Charge Compression Ignition (PCCI) engines, Reactive Control Compression Ignition (RCCI) engines, Gasoline Compression Ignition (GCI) engines, and Homogeneous Charge Compression Ignition (HCCI) engines.

Both spark ignition and compression ignition engines require an emission control system that includes one or more aftertreatment elements to limit the emission of undesirable pollutants, which are byproducts of combustion.

Spark ignition engines generally use a 3-way catalyst that both oxidizes unburned hydrocarbons and carbon monoxide and reduces Nitrogen Oxides (NO)_x). These catalysts require that the engine combustion be at or near a stoichiometric air/fuel ratio on average, so that it is possibleBoth oxidation and reduction reactions occur.

Since compression ignition engines generally run lean, they cannot rely on conventional 3-way catalysts to meet emission regulations. Instead, they use another type of aftertreatment device to reduce NO_xAnd (5) discharging. These aftertreatment devices may use a catalyst, lean in NO_xTraps and Selective Catalytic Reduction (SCR) reduce nitrogen oxides to molecular nitrogen. Additionally, diesel engines often require particulate filters to reduce soot emissions.

The fuel efficiency of an internal combustion engine can be substantially improved by varying the engine displacement. This allows maximum torque to be available when required, and also significantly reduces pumping losses and improves thermal efficiency by using smaller displacements when maximum torque is not required.

The most common method of implementing variable displacement engines today is to deactivate a group of cylinders substantially simultaneously. The commercially available variable displacement engines available today typically support only two or at most three displacements.

Another engine control method that varies the effective displacement of the engine is referred to as "skip fire" engine control. In general, skip fire engine control contemplates selectively skipping firing of certain cylinders during selected firing opportunities. Thus, a particular cylinder may be fired during one engine cycle, then skipped during the next engine cycle and selectively skipped or fired during the next engine cycle. From an engine cycle perspective, skip fire control may cause different groups of cylinders to be fired during successive engine cycles to produce the same average torque, while variable displacement operation deactivates the same groups of cylinders. In this way, even finer control of the effective displacement of the engine is possible. For example, firing every third cylinder in a 4-cylinder engine effectively reduces the engine maximum displacement to 1/3, which is a fractional displacement that cannot be achieved by simply deactivating a group of cylinders.

For skip fire engine control, selecting the correct firing density is a challenge. Using low firing densities to meet a given torque request is often fuel efficient; however, the engine may produce unacceptable levels of noise, vibration, and harshness (NVH). On the other hand, if the ignition density is higher than necessary, the engine efficiency may be reduced. Also, for compression ignition engines, the pollutants in the exhaust stream of the engine depend on the air-to-fuel ratio, which is affected by the firing density. Ignition density also affects the air flow through the engine, which may affect the efficacy of the aftertreatment system for reducing tailpipe emissions.

The delay associated with adjusting the ratio of fresh air to recirculated exhaust gas in the intake manifold and the intake manifold pressure also complicates controlling the ignition density through ignition density transitions. Accordingly, there is a need for more sophisticated methods for transitioning between different firing densities that provide high levels of engine efficiency, acceptable levels of NVH, and low levels of exhaust stream pollutants.

Disclosure of Invention

The invention relates to an engine controller arranged to operate a lean burn internal combustion engine having a plurality of cylinders in a skip fire mode. During operation, the engine controller is often required to transition from a first firing fraction to a target firing fraction to meet the changing torque request. During such transitions, the engine controller must handle several conflicting requirements, including meeting the requested torque request, preventing excessive noise, vibration and harshness (NVH), maintaining a proper air-fuel ratio, and controlling air flow through the internal combustion engine to avoid adversely affecting emissions.

For the present invention, the engine controller determines (a) the firing density or mode and (b) the pumping density or mode separately during the transition. For a defined firing density or pattern, NVH may be reduced while the torque request is satisfied. For a defined pumping density or pattern, the rate of air and/or recirculated exhaust gas pumped through the internal combustion engine is controlled, providing the ability to control air-to-fuel ratio and emissions. Thus, by blending ignition and deactivation of cylinders with pumping of unfueled cylinders during transitions, these conflicting requirements of meeting torque demand, NVH, maintaining a proper air-fuel ratio, and a desired air flow through the engine can all be balanced, resulting in more optimal results for each demand.

In various non-exclusive embodiments, the timing of initiating and/or completing the firing density transition relative to the pumping density transition may be further controlled to meet specific operating conditions. For example:

1. when the first firing density is greater than the target firing density, the engine controller is arranged to transition the firing density from the first firing density to the target firing density faster (i.e., in a shorter time) than the transition pumping density. By transitioning firing density faster, NVH is reduced while torque requirements are met. By extending the transition in pumping density, the inherent time delays associated with changes in manifold intake pressure and exhaust gas recirculation rate are better matched by changes in air flow rate of the engine, allowing for better control of the air-fuel ratio during the transition; or:

2. when the first firing density is less than the target firing density, the engine controller (i) first implements the change in pumping density and (ii) delays implementing the change in firing density until after the pumping density transition has begun. Again, this results in better control of the air-fuel ratio during the transition.

In still other embodiments, each firing opportunity during a firing density transition may have an associated action consisting of any of: firing a cylinder associated with the firing opportunity, thereby causing the cylinder associated with the firing opportunity to pump air through the engine; or to deactivate cylinders associated with the firing timing. In some embodiments, a first sigma-delta converter and a second sigma-delta converter may be used to select an action associated with any firing opportunity. In some embodiments, the first sigma-delta converter and the second sigma-delta converter generate a Fire-Enable (Fire-Enable) flag and a pumping-Enable (Pump-Enable) flag, respectively, for each firing opportunity of a cylinder of the internal combustion engine. If the ignition timing sets the pumping enable flag but not the ignition enable flag, the cylinder associated with the ignition timing will pump air through the engine without firing the cylinder.

In still other embodiments, the internal combustion engine is any type of lean-burn internal combustion engine, including, but not limited to, a compression-ignition engine and/or a diesel engine. In some embodiments, the engine is a lean burn engine fueled by gasoline.

Drawings

The invention, together with its advantages, may best be understood by reference to the following description taken in conjunction with the accompanying drawings in which:

FIG. 1A is a schematic illustration of a representative engine exhaust system for an exemplary compression-ignition engine.

FIG. 1B is a schematic illustration of an alternative representative engine exhaust system for an exemplary internal combustion engine.

FIG. 2 is a graph of exhaust gas temperature versus engine load for an exemplary internal combustion engine.

FIG. 3 is a schematic diagram of an engine controller for an exemplary internal combustion engine, according to a non-exclusive embodiment of the present disclosure.

Fig. 4 is a logic diagram of a skip fire controller arranged to operate an exemplary internal combustion engine, in accordance with a non-exclusive embodiment of the present invention.

FIG. 5 is a graph depicting the relationship between local fuel-air equivalence ratio and local combustion temperature for a representative compression-ignition engine.

Fig. 6A-6C are various graphs illustrating how a spike in particulate emissions of an exemplary internal combustion engine is mitigated during a transition from high ignition density to low ignition density, according to a non-exclusive embodiment of the present invention.

7A-7C are diagrams illustrating how NO of an exemplary internal combustion engine may be mitigated during a transition from low ignition density to high ignition density, according to non-exclusive embodiments of the present disclosure_xVarious plots of the surge in emissions.

FIG. 8 is a schematic block diagram of a valve deactivation controller according to a non-exclusive embodiment of the present invention.

FIG. 9 is a graph illustrating adjustment of exhaust gas temperature by using cylinder deactivation during skip fire operation of an exemplary internal combustion engine, according to a non-exclusive embodiment of the present disclosure.

FIG. 10 is a logic flow diagram illustrating prioritization of inputs provided to a valve deactivation controller in accordance with a non-exclusive embodiment of the present invention.

Fig. 11 is a flowchart illustrating steps for operating a lean-burn internal combustion engine having separately defined ignition density signals and pumping density signals during a transition from a first ignition density to a target point ignition density according to another non-exclusive embodiment of the present invention.

Fig. 12A illustrates transitions in ignition density and pumping density associated with transitioning a lean-burn internal combustion engine from a high ignition density to a lower ignition density, in accordance with the present disclosure.

Fig. 12B illustrates an exemplary ignition/pumping/deactivation pattern when transitioning a lean-burn internal combustion engine from a high ignition density to a lower ignition density, in accordance with the present disclosure.

Fig. 13A illustrates transitions in ignition density and pumping density associated with transitioning a lean-burn internal combustion engine from a low ignition density to a higher ignition density, in accordance with the present disclosure.

Fig. 13B illustrates an exemplary ignition/pumping/deactivation pattern when transitioning a lean-burn internal combustion engine from a low ignition density to a higher ignition density, in accordance with the present invention.

Fig. 14 is a logic diagram of a pair of sigma-delta converters for generating an ignition flag and a pumping flag for an ignition timing of a lean-burn internal combustion engine, according to a non-exclusive embodiment of the present invention.

FIG. 15 is a graph of normalized ratio fuel consumption versus engine torque for a gasoline-fueled engine using a lean air-fuel ratio and a stoichiometric air-fuel ratio.

Fig. 16 is a graph illustrating simultaneous ignition density and air-fuel ratio transitions in an exemplary internal combustion engine, according to a non-exclusive embodiment of the present disclosure.

In the drawings, like reference numerals are sometimes used to designate like structural elements. It should also be understood that the depictions in the figures are diagrammatic and not to scale.

Detailed Description

The invention relates to skip fire control that relies on a combination of torque request, exhaust temperature, NVH level, and air-fuel ratio to determine Firing Density (FD) of an internal combustion engine. The internal combustion engine may be a lean burn engine and may be used to power a vehicle. Additionally, the present invention also relates to the use of different types of working chamber valve control during skipped firing opportunities that allow or prevent pumping air through the engine to control and adjust exhaust gas temperature in the aftertreatment system.

Skip fire engine control

Skip fire engine control contemplates selectively skipping firing of certain cylinders during selected firing opportunities. Thus, for a given engine effective displacement that is less than the engine maximum displacement, a particular cylinder may be fired successively during one firing opportunity, skipped during the next firing opportunity, and then selectively skipped or fired during the next firing opportunity. The firing sequence may be expressed as a firing density, which indicates a ratio of firing opportunities for a firing to a total firing opportunity. The firing density may be expressed as a fraction, percentage, or in some other manner. For skip fire, much finer engine control is possible than engine control by using only a fixed set of deactivated cylinders.

A problem with skip fire controlled engines is the potential for unacceptable NVH associated with some firing sequences and cylinder loads. One way to deal with this problem is to not use a particular firing density or firing sequence that is known to produce unacceptable levels of NVH. Instead, other firing densities or firing sequences are used, and cylinder output is adjusted accordingly (e.g., by adjusting the mass of fuel injected into the cylinder) such that the desired engine output is delivered. Various methods of this type are described in commonly assigned U.S. patent application serial No. 13/654,244, which is incorporated herein in its entirety for all purposes.

During normal driving, the engine must typically operate over a wide range of engine speeds and engine loads. To meet these changing operating conditions, skip fire controlled engines may transition between various firing densities. For example, commercially available skip fire controllers are available that provide seventeen (17) different firing densities, each indicating a different reduced engine effective displacement. In contrast, for conventional variable displacement, one group of one or more cylinders is continuously fired, while a second group of one or more different cylinders is continuously deactivated or skipped. For example, an 8-cylinder conventional variable displacement engine may deactivate a cylinder bank (i.e., 2,4, or 6 cylinders) such that it operates using only the remaining (i.e., 6, 4, or 2) cylinders. With significantly more firing density available, skip firing provides significantly more detailed engine control than conventional variable displacement engine control.

The skip fire control generally operates based on firing timing, regardless of any particular cylinder firing pattern. An internal combustion engine having a plurality of cylinders is operated with a predefined sequence of firing timings for each engine cycle. For a 6-cylinder engine, the sequence of cylinder firing opportunities may be, for example, 1, 5,3, 6, 2, and 4, where the numbers 1 through 6 correspond to the physical location of each cylinder in the engine. In skip fire control, a skip fire mode or firing sequence may begin or end on any particular cylinder.

Dynamic Skip Fire (DSF) engine control

For some embodiments of skip fire engine control, the decision to fire or not fire a given cylinder of the engine is made dynamically, meaning on a firing opportunity by firing opportunity basis. In other words, prior to each successive firing opportunity, a decision is made to either fire or skip a firing opportunity. In various embodiments, the firing sequence is determined on a firing opportunity by firing opportunity basis by using a sigma-delta converter or, equivalently, a delta sigma converter. Such a skip fire control system may be defined as dynamic skip fire control. For more details on DSFs, see U.S. patent nos. 7,849,835, 9,086,020, and 9,200,575 and U.S. application No. 14/638,908, each of which is incorporated by reference herein for all purposes.

Skip fire engine control (including DSF) may provide various advantages, including substantially improving fuel economy of the spark-ignition engine, wherein pumping losses may be reduced by operating at higher average MAP levels. For compression ignition engines, skip fire control provides a means for controlling engine exhaust temperature over a wide range of engine operating conditions. In particular, skip fire control may be used to adjust exhaust temperature in a range where the aftertreatment emission control system may efficiently reduce tailpipe emissions. Various methods of this type are described in commonly assigned U.S. patent application serial No. 15/347,562, which is incorporated herein in its entirety for all purposes. The use of skip fire control may also provide fuel consumption efficiency improvements of over twenty percent (20%) in compression ignition engines at light loads (e.g., loads less than 10% of the engine's maximum output).

Exemplary lean burn Engine and exhaust System

FIG. 1A is a schematic illustration of an exemplary system 10 that includes an engine 12 and an aftertreatment system 14A. In this embodiment, the particulate filter 20 is located upstream of the reduction catalytic converter 26 and the oxidation catalytic converter 22. This arrangement may be particularly suitable for gasoline-fueled lean burn engines.

The engine 12 includes a plurality of cylinders 16 in which combustion occurs. In the illustrated embodiment, the engine 12 includes four (4) cylinders 16. It should be appreciated that the engine 12 as illustrated is merely exemplary, and that fewer or more cylinders 16 may be included. Additionally, the engine 12 may be a compression ignition engine, a Spark Ignition (SI) engine, an engine that combines spark ignition with compression ignition, or an engine that ignites air and fuel mixtures using different techniques. For simplicity, the following discussion of the operation of the engine 12 is primarily in the context of a compression ignition engine (such as a diesel engine). However, it should be understood that many of the features discussed below are equally applicable to other types of engines, such as SI engines.

Exhaust system 14A may additionally include one or more temperature sensors. Such temperature sensors may include (a) a temperature sensor 34 for monitoring the temperature of the particulate filter 20, (b) a temperature sensor 36 for monitoring the temperature of the oxidation catalytic converter 22, and (c) a temperature sensor 38 for monitoring the temperature of the reduction catalytic converter 26.

Referring to FIG. 1B, a system 10 is shown that includes a lean burn engine 12 and an alternative representative exhaust system 14B. With this arrangement, the particulate filter 20 is placed downstream of the reduction catalyst 26. Otherwise, exhaust systems 14A and 14B are substantially identical.

The arrangement of exhaust system 14B may be advantageous when particulate filter 20 needs to be cleaned periodically by an active process that raises its temperature to burn off accumulated soot. Those temperatures typically reach 500C to 600C. The active cleaning process may include: unburned hydrocarbons are intentionally introduced into the exhaust stream and oxidized in the oxidation catalytic converter 22 to generate heat. By positioning the oxidation catalyst 22 upstream of the particulate filter 20, the temperature within the particulate filter 20 may be actively controlled during the cleaning process.

It should be noted that the particular order of the various post-processing elements shown in fig. 1A and 1B is merely exemplary and should not be construed as limiting. The order of the various post-processing elements described herein, as well as additional post-processing elements that may be used, may be widely varied to meet operating conditions, regulatory requirements, and/or other objectives.

It should also be noted that the exhaust systems 14A and 14B may include other types of sensors in addition to temperature sensors. Such sensors may include (not shown), for example, oxygen sensors placed before and after the oxidation catalytic converter 22, and NO downstream of the reduction catalytic converter 26_xA sensor.

It should be further noted that various other features and elements not shown in fig. 1A and 1B may be located between the engine and the aftertreatment elements of exhaust systems 14A and 14B. Such elements may include, but are not limited to, an exhaust gas recirculation system (EGR), a turbine for powering a turbocharger, a wastegate to control the flow of exhaust gas through the turbine, and the like.

Exhaust system operating temperature

As the exhaust flow passes from the engine 12 through the exhaust manifold 18, the exhaust flow will generally be at its hottest temperature. As the exhaust flow passes through subsequent elements of the exhaust system 14A/14B, the gases tend to cool from one stage to the next. Thus, aftertreatment elements 20, 22 and 26 are generally arranged in order of highest to lowest operating temperatures as desired. For example, the exhaust passing through the particulate filter 20 is hotter than the exhaust passing to the downstream elements 20 and 26 in FIG. 1A. In the arrangement of fig. 1B, the gas passing through the reduction catalyst 26 is hotter than the gas of the downstream particulate filter 20. It should be appreciated that the exothermic chemical reaction may occur in any aftertreatment component, which may raise the temperature of the aftertreatment component and any other downstream components.

In order for aftertreatment systems 14A and 14B to function properly, elements 20, 22, and 26 each need to operate within a specified elevated temperature range. In a non-exclusive example, a representative operating range for the reduction catalyst 26 is in the approximate range of 200 ℃ to 400 ℃. It should be understood that these temperature values are approximate rather than absolute. For example, each temperature value may vary within ten percent (+/-10%) of 200 ℃ and 400 ℃. If reduction catalyst 26 (including reduction catalysts 28, 30) were last for a given aftertreatment system, the upstream elements, including particulate filter 20 and oxidation catalyst 22, regardless of their order, would typically operate at a slightly higher temperature range.

Referring to FIG. 2, a graph 40 is illustrated depicting exhaust gas temperature versus operating load at the exhaust manifold 18 of a representative boosted compression-ignition engine 12 operating at 1250rpm without cylinder deactivation.

In this example, curve 42 represents exhaust temperature as a function of engine load, expressed as Brake Mean Effective Pressure (BMEP), for all engine cylinders fired under substantially the same conditions.

The operating range 44 is a temperature range of the exhaust gas in the exhaust manifold 18 that results in efficient operation of the aftertreatment system. In this particular example, the operating range is approximately 225 ° to 425 ℃.

As previously described, the exhaust gas will typically cool slightly at each stage of any aftertreatment system 14A/14B. For example, when the exhaust gas reaches the reduction catalytic converters 28,30 in the aftertreatment system 14A, the temperature may have dropped by approximately 25 ℃. In other words, the temperature of the exhaust gas is at or near a representative operating range of the reduction catalytic converter 26, which, as noted above, may be 200 ℃ to 400 ℃.

It is important to note that the temperature ranges provided at the exhaust manifold 18 and at the final stages of the aftertreatment system 14A/14B are merely exemplary, and should not be construed as limiting in any way. Conversely, different engine operating points and engine designs may have different starting, intermediate, and ending temperatures and temperature offsets between the exhaust manifold 18 and the final element of the aftertreatment system 14A/14B. In fact, in some cases, the exhaust temperature may rise in the exhaust system due to exothermic chemical reactions. Thus, actual temperature values and ranges as provided herein should not be construed as limiting the scope of the invention.

Examination of fig. 2 indicates that: a substantial portion of the engine 12 operating range falls outside the preferred steady state operating range (i.e., at available NO)_xOutside the acceptable range of removal). Advantageously, as described in more detail below, skip fire engine control may be effectively used as a strategy for adjusting and maintaining exhaust gas temperature within a preferred steady state operating range.

Skip fire control system

Referring to FIG. 3, a schematic diagram of an engine controller 50 is illustrated that illustrates several controls and/or systems for controlling operation of the engine 12 in a skip fire mode. These control systems include an Exhaust Gas Recirculation (EGR) system controller 52, a turbocharger controller 54, a fuel control unit 56, and a skip fire controller 58.

EGR system controller 52 operates to recirculate a portion of the exhaust gas back to cylinders 16 of engine 12. Recirculation tends to dilute the fresh air intake stream entering the cylinders with gases inert to combustion. The exhaust gas acts as an absorber of the heat generated by combustion and reduces peak temperatures within the cylinders 16. As a result, NO is generally reduced_xAnd (5) discharging. For example, in a compression ignition diesel engine, exhaust gas replaces a portion of the oxygen in the pre-combustion mixture. Since NO is mainly formed when a mixture of nitrogen and oxygen is subjected to high temperatures_xThus, lower combustion temperatures and reduced amounts of oxygen in the working chamber cause the generation of NO_xA reduction in the amount of (c).

The boost controller 54 controls the amount of compressed air directed into the cylinders 16 of the engine 12. Boosting (i.e., supplying compressed air to the engine 12) allows more power to be generated than in naturally aspirated engines, as more air and proportionally more fuel may be input into the cylinders 16. Boost controller 54 may operate with a turbocharger, a super booster, or a dual booster. An important difference between turbochargers and superchargers is that superchargers are mechanically driven by the engine (often through a belt connected to the crankshaft), while turbochargers are powered by a turbine driven by the engine exhaust. Turbochargers tend to be more efficient, but less responsive, than mechanically driven superchargers. Twin boosters refer to engines having both super-and turbo-chargers.

Fuel control unit 56 is used to determine the amount of fuel required for cylinders 16 of engine 12. The amount of fuel injected is primarily based on the torque request because for lean-burn engines, the efficiency of torque production is not strongly affected by the air/fuel ratio. However, there must be sufficient air flow into the engine to combust the delivered fuel and result in lean burn operation. Most vehicles rely on a mass air flow sensor to determine air mass. Given the mass of air flow into the engine and the mass of fuel injected, the air-to-fuel ratio may be determined as one of the inputs to the engine controller 50. Based in part on this value, fuel control unit 56 determines how much fuel to inject into cylinders 16 of engine 12. As previously mentioned, the air-to-fuel ratio of a diesel engine may range from approximately 16 to 55 compared to a stoichiometric air-to-fuel ratio of 14.6.

The skip fire controller 58 is responsible for determining whether the engine 12 should be operated in either the maximum displacement mode or the skip fire mode. When no firing fraction other than 1 adequately meets the high torque demand, then the skip fire controller will operate the engine 12 at maximum displacement. Otherwise, in skip fire mode, the engine is typically operated at one of a plurality of reduced effective displacements, each defined by a different firing density or fraction.

When in skip fire mode, skip fire controller 58 is responsible for determining the firing density or firing fraction that meets the current torque request. In other words, skip fire controller 58 defines a firing fraction suitable for (1) satisfying the current torque request and (2) operating the vehicle at an acceptable NVH level. Meeting these two constraints has the highest priority in the engine control architecture as a whole. Other parameters that may also be optimized are (3) fuel efficiency, (4) exhaust temperature, and (5) air/fuel ratio. The point (3) need not be explained because it is obviously advantageous to minimize the fuel consumption. Points (4) and (5) are the desire to reduce the burden on aftertreatment elements in the exhaust system and improve exhaust pipe emissions. As driving conditions change (i.e., engine speed and torque demand change), the skip fire controller 58 is responsible for selecting different firing fractions, each indicating a different reduced effective displacement less than the maximum displacement of the engine 12, best meeting the five targets (1) through (5) expressed above in the specification.

The skip fire controller 58 receives at least three inputs, including (a) a current torque request, (B) a signal 60 indicative of the temperature of the exhaust gas in the aftertreatment system 14A/14B, and (c) the air-to-fuel ratio of the active cylinder or cylinders 16 of the engine 12. In response, skip fire controller 58 generates firing density or fraction 62. With these three inputs, skip fire controller 58 can provide even finer control of engine 12, selecting the optimal firing density that best meets the above-mentioned targets (1-5).

The turbocharger controller 54 receives (a) the current torque request, (B) a signal 60 indicative of the temperature of the exhaust gas in the aftertreatment system 14A/14B, (c) the air-to-fuel ratio provided to one or more of the cylinders 16 of the engine 12, and (d) the output of the skip fire controller 58. In response, the boost controller 54 determines the amount of compressed air to be introduced into the cylinders 16 of the engine 12.

The EGR system controller 52 similarly receives (a) the current torque request, (B) a signal 60 indicative of the temperature of the exhaust gas in the aftertreatment system 14A/14B, (c) the air-to-fuel ratio provided to one or more of the cylinders 16 of the engine 12, and (d) the output of the skip fire controller 58. In response, the EGR system controller 52 determines an amount or percentage of exhaust gas to be recirculated. Again, by receiving three inputs (a), (b), and (c), the EGR system is able to more accurately determine the amount of exhaust gas to be recirculated.

The outputs of the skip fire controller 58, boost controller 54, and EGR controller 52 are then all considered to produce an air intake value 68 that is provided to the cylinders 16 of the engine 12. Additionally, as described above, the fuel control unit 56 takes into account the air intake value 68 when providing the appropriate amount of fuel to the cylinders 16 of the engine 12. The fuel and air together define an air-fuel mixture that is provided to the cylinders 16, which is characterized by an air-fuel ratio.

In a non-exclusive embodiment, skip fire controller 58 is a dynamic skip fire controller. In other words, skip fire controller 58 dynamically makes a decision to fire or misfire a given cylinder of the engine, meaning on a firing opportunity by firing opportunity basis.

Referring to fig. 4, a logic diagram of skip fire controller 58 is illustrated.

Skip fire controller 58 includes a firing fraction calculator 70, a firing timing determination unit 74, a powertrain parameter adjustment module 76, a firing control unit 78, and an aftertreatment monitor 80.

The firing fraction calculator 70 may receive at least three inputs including (a) the current torque request 72A, (B) the exhaust temperature as provided by the aftertreatment monitor 80 receiving the signal 60, and (c) the target air-to-fuel ratio 72B. In response, firing fraction calculator 70 determines the skip fire firing fraction or firing density that best matches the above-described target. It should be appreciated that the firing fraction or density may be communicated or represented in a variety of ways. For example, the firing fraction or firing density may take the form of a firing pattern, sequence, or any other firing characteristic that relates to or inherently conveys the firing percentage or density described above.

In still other embodiments, the firing fraction calculator 70 may take into account other information in determining the firing density. Such other information may include, for example, vehicle speed, engine speed, transmission gear ratio, oxygen sensor data, NO_xSensor data, ambient air temperature, exhaust gas temperature, catalyst temperature, barometric pressure, ambient humidity, engine coolant temperature, and the like. In various embodiments, as these parameters change over time, the firing fraction may be dynamically adjusted in response to these changes.

Aftertreatment monitor 80 represents any suitable module, mechanism, and/or sensor(s) that obtains data related to the temperature of an aftertreatment element. If the reduction catalyst 26 has the narrowest operating range of any aftertreatment element, then data representative of its temperature may only be used. Alternatively, the aftertreatment temperature may correspond to the temperature of any or all of the particulate filter 20, the oxidation catalytic converter 22, and/or the reduction catalytic converter 26 (see fig. 1A and 1B). In various embodiments, for example, aftertreatment monitor 80 may include or work in cooperation with oxygen sensor data from an oxygen sensor in aftertreatment system 14A/14B, and N placed before and after reduction catalytic converter 26O_xA sensor. Aftertreatment monitors 80 may also include inputs such as ambient air temperature, exhaust gas temperature in the exhaust manifold, barometric pressure, ambient humidity, and/or engine coolant temperature.

In some embodiments, skip fire controller 58 and aftertreatment monitor 80 do not require direct measurement or sensing of the temperature of the aftertreatment element. Instead, an algorithm using one or more inputs (such as a catalytic converter temperature model) may be used to estimate aftertreatment element or system temperature. The model may be based on the above parameters representative or related to the catalytic converter temperature (e.g., oxygen sensor data, NO)_xSensor data, exhaust temperature, ambient temperature, barometric pressure, ambient humidity, etc.).

The ignition timing determination unit 74 receives input from the firing fraction calculator 70 and/or the powertrain parameter adjustment module 76, and is arranged to issue a series of firing commands (e.g., drive pulse signals) that are provided to a firing control unit 78. The ignition timing determining unit 74 may take a variety of different forms. For example, in some embodiments, the spark timing determination unit 74 may utilize various types of look-up tables to implement the desired control algorithm. In other embodiments, a sigma-delta converter or other mechanism is used. A series of firing commands (sometimes referred to as drive pulse signals 75) are provided to a firing control unit 78 that orchestrates the actual firing of the cylinders 16 of the engine 12.

The powertrain parameter adjustment module 76 directs the ignition control unit 78 to appropriately set the selected powertrain parameters to ensure that the actual engine output is substantially equal to the requested engine output at the commanded ignition fraction or density. For example, the powertrain parameter adjustment module 76 may be responsible for determining a desired fueling level, a number of fuel injection events, a fuel injection timing, an Exhaust Gas Recirculation (EGR) level, and/or other engine settings desired to help ensure that the actual engine output matches the requested engine output.

The ignition control unit 78 receives input from the ignition timing determination unit 72 and the powertrain parameter adjustment module 76. Based on the inputs, the ignition control unit 78 directs the engine to operate in the ignition sequence 75 determined by the ignition timing determination unit 74 using the engine parameters determined by the powertrain parameter adjustment module 76.

For example, some suitable firing fraction calculators, firing timing determination units, powertrain parameter adjustment modules, and other associated modules are described in the commonly assigned ones of: U.S. patent nos. 7,954,474; 7,886,715, respectively; 7,849,835, respectively; 7,577,511, respectively; 8,099,224, respectively; 8,131,445, respectively; 8,131,447, respectively; 9,086,020, respectively; and 9,120,478; U.S. patent application No. 13/774,134; 13/963,686, respectively; 13/953,615, respectively; 13/886,107, respectively; 13/963,759, respectively; 13/963,819, respectively; 13/961,701, respectively; 13/843,567, respectively; 13/794,157, respectively; 13/842,234, respectively; 13/004,839, respectively; 13/654,244, and 13/004,844, each of which is incorporated by reference herein in its entirety for all purposes.

Air-fuel ratio range for compression ignition engines

Compression ignition engines can operate over a wide range of air to fuel ratios. For example, a diesel engine may operate at an air-fuel ratio ranging from 16 to 55 depending on speed/load conditions. As previously mentioned, the torque of a diesel engine is determined by the amount of fuel injected rather than the mass of air directed because there is excess oxygen to combust all of the fuel. However, the inducted ingested air mass affects the air-fuel ratio, which in turn affects exhaust gas temperature and engine emissions.

Referring to FIG. 5, a graph depicting the relationship between local fuel/air equivalence ratio and combustion temperature for a representative diesel engine is illustrated. As depicted, the local fuel-air equivalence ratio is provided along the vertical axis, while the temperature in degrees Kelvin, or "K", of the local temperature during combustion is provided along the horizontal axis. The fuel-air equivalence ratio is defined as the ratio of the fuel-to-oxidant ratio to the stoichiometric fuel-to-oxidant ratio.

The air-fuel ratio provided along the vertical axis is considered "local" because diesel engines are typically stratified, meaning that the air-fuel mixture is heterogeneous within the cylinder. As a result, the concentration of fuel is higher in some regions (e.g., generally near the top of the cylinder where fuel is injected) than in other regions of the cylinder (e.g., near the bottom, away from the injected fuel jet). The fuel is injected in a plurality of jets which are dispersed into fine droplets, and combustion occurs mainly in the vicinity of the fuel injection jets.

Diesel engine operation generally occurs within the area defined by the oval 30. The operation in the upper left side of the ellipse 30 corresponds to operation at an average air/fuel ratio close to stoichiometry. The operation in the lower right side of the ellipse 30 corresponds to the operation at a high average air-fuel ratio. Note that due to diesel engine combustion non-uniformity, the average air/fuel ratio is not simply the inverse of the fuel air equivalence ratio multiplied by the stoichiometric air/fuel ratio (14.6).

The horizontal dashed line represents a stoichiometric or 1 fuel-to-air equivalence ratio (corresponding to an air-to-fuel ratio of approximately 14.6 as previously described). Above stoichiometry, the fuel-air ratio is rich. Below the stoichiometric ratio, the fuel-air ratio is lean. For reference, typical operating regions of a Spark Ignition (SI) engine are also labeled. SI engines in NO_xOperation in the production zone, meaning that some type of NO is generally required_xA reduction aftertreatment element (such as a three-way catalyst).

The graph further shows the soot region defined by the "soot formation limit" line. The operating region within the soot formation limit line produces a detectable amount of soot, with more soot being produced as the shadow darkens. When operating in the soot region, a particulate filter is typically required to reduce soot in the exhaust stream prior to releasing the exhaust stream into the atmosphere. In order to minimize the cleaning burden on the particulate filter, operation in areas with darker shading should be avoided or at least minimized.

The graph further shows NO_xA discharge area. Again, the darker shaded areas indicate undesirable pollutants in the exhaust stream (NO in this case)_x) Is/are as followsThe concentration is higher. Reduction catalytic converters are often required to reduce NO_xAnd (5) discharging. As with soot, desirably away from NO_xOperating in the region of highest emissions to reduce NO_xThe burden of the post-processing system. Operation in the dashed area 40 is desirable from an emissions standpoint, as soot and NO_xBoth emissions are inherently low without any post-treatment. An air-to-fuel ratio between 20 and 35 may be desirable so that most combustion may occur in operating region 40. Operation within region 40 may be maintained by adjusting spark density, boost level, exhaust gas recirculation level, and lift profiles of intake and exhaust valves. In some embodiments, the engine may be controlled in a manner that meets environmental regulations without the use of any aftertreatment systems. Alternatively, the size and complexity of the aftertreatment system may be reduced, as the pollutant load that the aftertreatment system needs to remove is reduced.

Thus, the graph of FIG. 5 demonstrates controlling the value of air-fuel ratio as part of a skip fire control algorithm, particularly for compression ignition engines. With the addition of air/fuel ratio and exhaust temperature as inputs to the skip fire controller 58, an appropriate firing density may be selected for the desired torque output, air/fuel ratio, and exhaust temperature. The selected firing density may or may not be optimized for fuel consumption. However, in some cases, achieving low tailpipe emissions to meet regulatory requirements is more important than absolute maximum fuel efficiency.

To accurately control the air-fuel ratio in the firing cylinder, the inducted Mass Air Charge (MAC) must be accurately estimated. As disclosed in U.S. patent No. 9,945,313 and U.S. patent application serial No. 15/628,309, which are incorporated by reference in their entirety for all purposes, determining the mass of air directed is more complex in a skip fire controlled engine than in an engine operating at a fixed displacement. Adjustments to cylinder MAC may be made prior to firing the cylinder based on cylinder firing history and engine skip fire sequences.

As described above, controlling the air-fuel ratio in the ignition cylinder is an important element in the skip fire control for a lean burn engine. Additionally, control of fuel injection timing and fuel injection mode is also important. Even if the air-fuel ratio is the same in different duty cycles, the torque produced and the combustion byproducts may differ depending on the timing and mode of fuel injection. Fuel injection timing refers to when fuel is injected in a cylinder relative to the position of the piston in the cylinder (usually indicated by crank angle). The fuel injection mode refers to the number and duration of discrete fuel injection events that may occur during the operating cycle. For example, where there are multiple injection events in a duty cycle (which collectively deliver the desired fuel mass), the fuel injection may be discontinued rather than injecting all of the fuel in a single event.

A cylinder skipped on one or more previous operating cycles will have cooler cylinder walls than if it had been fired. The greatest effect on cylinder wall temperature comes from the immediately preceding duty cycle, but the cylinder firing history of approximately the past five firing occasions will affect cylinder wall temperature. Additionally, if the previous firing opportunity was skipped instead of firing, the composition of the residual gas in the cylinder may be different. These changes in initial conditions at the start of combustion can affect combustion dynamics. To adjust for different initial combustion conditions, the timing and pattern of fuel injection may be optimized based on cylinder firing history. The desired air-fuel ratio may also be optimized. Adjusting the fuel injection timing, fuel injection mode, and air-fuel ratio may reduce undesirable combustion products and improve fuel efficiency in the fired cylinder.

Additionally, using three inputs as described herein provides flexibility and control to optimize air-fuel ratio and/or exhaust temperature in this manner during transitions from one steady state to another steady state condition (i.e., transitions from one ignition density to another ignition density) to minimize spikes in engine emissions during ignition density transitions, which was not previously possible.

Reducing emissions during ignition density transitions

FIG. 6A illustrates a representative engine 12 including four cylinders 16. Engine 12 is arranged to receive fuel from fuel control unit 56 (see fig. 3), and intake airflow from a combination of EGR system 88 controlled by EGR system controller 52 (see fig. 3) and turbocharger system 90 controlled by boost controller 54 (see fig. 3). EGR system 88 includes an EGR flow valve that controls the flow of exhaust gas back into the intake airflow of engine 12. Turbocharger system 90 includes an exhaust turbine 92, a shaft 94, and a compressor wheel 96. Compressor wheel 96 is the portion of the compressor that is used to increase the pressure in the intake manifold above atmospheric pressure. An optional cooler 98 may also be provided to cool the intake air, allowing for a higher MAC. Air from the intake manifold is directed into the cylinders through one or more intake valves on each cylinder.

Fig. 6A depicts an example of a Firing Density (FD) transition from 1.0 to 0.5. Prior to the transition, all four cylinders are fired, as illustrated on the left side of the figure. After the transition, only two of the four cylinders are fired, while the remaining two cylinders are skipped, as graphically illustrated by an "X" passing through two of the four cylinders 16 on the right side of the figure.

Fig. 6B-6C are graphs depicting the behavior of various engine parameters through ignition density transitions, such as the transition from an ignition density of 1 to an ignition density of 0.5 depicted in fig. 6A. Fig. 6B depicts engine parameter behavior during abrupt ignition density transitions, and fig. 6C depicts engine parameter behavior during gradual or smooth ignition density transitions.

A sudden transition may refer to a situation where the engine is firing density transitioning in one engine cycle or 4 firing occasions for a 4 cylinder engine, which requires two engine revolutions for a 4 stroke engine. For an engine operating at 1500rpm (typical engine speed), a sudden transition occurs within a period of 80 milliseconds. As shown in fig. 6B, when the ignition density is directly shifted from 1.0 to 0.5, the fuel flow rate per operating cylinder is immediately increased, so that the torque demand can now be satisfied by 2 cylinders instead of 4 cylinders (assuming constant during the shift). However, due to the inherent lag in turbocharger system 90 adjusting intake manifold pressure for the new firing density, the air flow per cylinder can only be adjusted gradually. As a result, the air-fuel ratio is greatly decreased during the transition, causing a temporary surge in particulate emissions (PM) during the transition.

In contrast, fig. 6C shows how PM surge can be significantly reduced by managing the transition from high ignition density to lower ignition density. Management of the transition involves firing at one or more intermediate firing densities between the higher firing density and the lower firing density. Equivalently, the transition between the initial firing density and the final target firing density may be considered to occur over multiple engine cycles. For example, if the initial firing density is 1.0 and the target firing density is 0.5, one or more of (0.9, 0.8, 0.7, and 0.6) intermediate firing densities may be used to smooth the transition. By using one or more intermediate firing densities, the air-to-fuel ratio in the fired cylinder 16 remains relatively stable and, as a result, PM spikes are minimized. Depending on the exact nature of the transition, the transition from the initial firing fraction to the target firing fraction may occur over approximately 3 to 15 engine cycles.

7A-7C are graphs illustrating how NO of an exemplary internal combustion engine is mitigated during a transition from low to high ignition density_xVarious plots of the surge in emissions. NO abatement_xThe surge in emissions is substantially complementary to the measures described above. When the engine 12 is operating at a relatively low firing density and transitioning to a higher firing density (e.g., a transition from 0.5 to 1.0), a gradual transition using one or more intermediate firing densities (e.g., 0.6, 0.7, 0.8, 0.9) tends to reduce NO_xThe emission is dramatically increased.

Fig. 7A is similar to fig. 6A, except that the engine 12 is initially at a low firing density of 0.5 and transitions to a high firing density of 1. Fig. 7B shows the behavior of various engine parameters during an abrupt transition from low to high ignition density. TheThe transition causes the fuel flow to each fired cylinder to immediately make a falling transition. However, due to the lag, the air intake gradually shifts lower. As a result, the air-fuel ratio in the ignited chamber is abruptly increased, thereby causing NO_xThere is a corresponding surge in emissions. In contrast, as shown in FIG. 7C, the controlled, gradual transition reduces or completely eliminates spikes in air-to-fuel ratio and resulting NO_xThe emission is dramatically increased.

Controlling firing density to regulate exhaust temperature

Skip fire controller 58 may be used to adjust the temperature of the exhaust gas in several ways. First, the temperature can be controlled by skipping over cylinder 16 or firing the cylinder. Second, for skipped cylinders, the cylinder is either pumped through or deactivated by closing one or both of the intake valve(s) or exhaust valve(s) so that no air is pumped through the cylinder.

The different firing densities vary the workload on each cylinder 16 of the engine 12. If many or all of the cylinders are fired, less work is performed per cylinder 16. If fewer cylinders are fired, more work is performed per cylinder 16. In general, the more work a given cylinder 16 performs, the higher the temperature of the exhaust gas from that cylinder.

As described above with respect to fig. 2, a representative operating range for reduction catalytic converter 26 may be in the approximate range of 200 ℃ to 400 ℃. By changing the ignition density, it is therefore possible to adjust the temperature of the exhaust gas to be substantially maintained within the range of 200 ℃ to 400 ℃. By maintaining the temperature within the desired temperature range, reduction catalyst 26 reduces NO_xMore effective in terms of emission.

Consider an example of a six cylinder engine having three active cylinders and three deactivated or skipped cylinders. Each active cylinder will receive a relatively rich air/fuel ratio. By activating the fourth cylinder (or cylinders), the same amount of fuel is spread out among the four cylinders (or cylinders). As a result, the air-fuel ratio of each cylinder becomes leaner. The leaner the air-fuel ratio, the cooler the resulting exhaust gas will be. Thus, by activating more cylinders, the temperature of the exhaust gas can be controlled and reduced. If too cold, a supplement to the above measures may also be used to increase the temperature of the exhaust gas. If five or six cylinders are operational and the exhaust is too cold, one or a few cylinders may be deactivated. By deactivating cylinders, the air-fuel ratio spread across the remaining active cylinders becomes richer. As a result, the combustion produces hotter exhaust gases.

Skipping of cylinders 16 may be implemented in one of two ways. First, either the intake valve or the exhaust valve (or both) may be closed during the skipped firing opportunity. As a result, no air is pumped through the cylinder. Second, both the intake and exhaust valves may be open, but no fuel is provided to the cylinder during the skipped firing timing. As a result, air is pumped through the chamber, but is not combusted. When air is pumped into the exhaust system, its function is to reduce the temperature of the exhaust gases. Thus, by allowing the skipped cylinders to pump or not pump air, the temperature of the exhaust gas may be further controlled or adjusted.

Referring to FIG. 8, a schematic block diagram of the valve deactivation controller 102 is shown. In various embodiments, valve activation controller 102 may be included in or separate from skip fire controller 58.

Valve deactivation controller 102 receives input 104 indicative of an actual and/or estimated value of exhaust gas temperature in aftertreatment system 14A/14B, input 106 indicative of an amount of compressed air forced into cylinders 16 by turbocharger 90 (or some other type of boosting system), a current torque request 108, and input 110 indicative of firing density as determined by skip fire controller 58. In response, valve deactivation controller 102 makes any of the following decisions for the skipped cylinders 16:

(1) preventing pumping by closing either of the intake and/or exhaust valves of the skipped cylinder 16; or

(2) Air is allowed to be pumped through the deactivated cylinders 16 by opening both the intake and exhaust valves. Since no fuel is provided to the cylinders 16, no combustion occurs, and intake air is pumped into the aftertreatment system 14A/14B.

FIG. 9 is a chart 114 illustrating how the exhaust temperature may be adjusted using the valve deactivation controller 102. Curve 115 represents the reduction catalyst efficiency as a function of the reduction catalyst temperature. As noted, a representative operating range for the reduction catalyst 26 may be in the range of 200 ℃ to 400 ℃, which is defined by the region between the dashed lines 117A and 117B. Line 116 represents a threshold temperature value below the maximum operating range of 400 deg.c. The exact temperature value of threshold 116 may vary, but in general, it represents a preferred or target operating temperature for reduction catalytic converter 26. The threshold may be shifted toward the upper bound of the operating temperature range, as shown in FIG. 9, because pumping air through the engine will rapidly decrease the catalyst temperature without pumping air having a more gradual effect on the catalyst temperature. When the actual exhaust temperature 104 is in region 118A (i.e., below threshold 116), valve deactivation controller 102 operates to prevent pumping of the skipped cylinders 16. As a result, the temperature in the aftertreatment system 14A/14B will be substantially maintained or prevented from decreasing. If the actual exhaust temperature 104 is in region 118B (i.e., above threshold 116), valve deactivation controller 102 operates to allow the skipped cylinder 16 to be pumped. As a result, the temperature in the aftertreatment system 14A/14B will tend to decrease.

Thus, the ability to control or adjust the temperature of the exhaust gas may be implemented by (1) firing the cylinder or skipping the cylinder, and/or (2) by allowing or preventing pumping of the skipped cylinder. In alternative embodiments, only cylinder firing/skipping may be used, or alternatively, both techniques may be used cooperatively.

Priority level

Most turbochargers rely on a contour efficiency map, which specifies a high efficiency region defined by (a) a specific pressure range and (b) an air volume range. The contour efficiency map also defines surge lines that should not be exceeded. If the pressure exceeds the surge line, the chances of mechanical damage to the turbocharger and/or the engine increase significantly.

Cylinder deactivation tends to increase MAC and pressure in the firing cylinders. For one or more deactivated cylinders, the intake air is shared among the less active cylinders, causing an increase in pressure compared to the pressure in the case where all cylinders are active. As a result, cylinder deactivation may increase the chance of exceeding the surge line of the turbocharger. Accordingly, a priority scheme is desired that balances priority between torque requests, exhaust gas temperature, and preventing the turbocharger from exceeding its surge line.

Referring to fig. 10, a logic flow diagram 120 illustrating a prioritization scheme implemented by skip fire controller 58 and/or valve deactivation controller 102 is shown. The intent of the prioritization scheme is to continuously monitor exhaust gas temperature, turbocharger operating conditions, and current torque requests and prioritize which should be precedent based on real-time operating conditions. Since the likelihood of turbocharger and/or engine damage increases dramatically in the event of a surge line being exceeded, it is preferable to set the turbocharger to the highest priority. The second priority is exhaust temperature, which may damage the exhaust system if it exceeds the operating limits for an extended period of time, or which may result in unacceptable emissions if the exhaust temperature is outside of its normal operating region. The third priority is a desire to meet the requested engine torque.

The logic flow diagram begins and proceeds to decision step 122. In decision step 122, the turbocharger is continuously monitored. If the turbocharger compressor is operating at or near the surge line, the turbocharger is given the highest priority and the flowchart proceeds to step 124.

In step 124, the skip fire controller 58 and/or the valve deactivation controller 102 operate to move the turbocharger operating point away from the surge line so that there is sufficient operating margin to avoid surge. Such actions may include activating more cylinders, reducing air intake, reducing pressure generated by the turbocharger, reducing EGR flow, reducing engine torque, and so forth.

On the other hand, if the turbocharger is operating sufficiently far from the surge line, the flow chart proceeds to decision step 126. In decision step 126, it is determined whether the exhaust is operating outside of a predefined normal range (e.g., 200 ℃ to 400 ℃).

If so, the exhaust temperature is set to the highest priority (step 128). Skip fire controller 58 and/or valve deactivation controller 102 operate to regulate exhaust gas temperature to within its normal range by activating or deactivating cylinders and/or reducing engine torque.

If the exhaust temperature is within its normal operating range, the current torque request is set to priority (step 130). As a result, the firing density defined by skip fire controller 58 is predominately defined as meeting the current torque request. The air-fuel ratio may be controlled such that the engine generally operates in region 40 of fig. 5, where engine-generated pollutants are minimized.

The above decision is made continuously during operation. As a result, the priority of skip fire controller 58 and/or valve deactivation controller 102 is continuously updated to meet the current operating conditions. When problems arise with the turbocharger and/or aftertreatment system, they are prioritized and corrected. When there is no problem, then it is a priority to meet the current torque demand.

Ignition density transition problem

During operation, skip fire controller 58 is often required to transition from the first firing fraction to the target firing fraction to meet the changing torque demand. During such transitions, skip fire controller 58 has several constraints, including meeting requested torque, preventing excessive NVH, controlling air flow through internal combustion engine 12, operating with minimal pollutant emissions, and maintaining a proper air-to-fuel ratio. In some cases, these constraints conflict with each other, and satisfying one constraint may cause another constraint to exceed some pre-established limit. For example, if too much air is pumped through the internal combustion engine 12, the exhaust gas temperature tends to decrease, which may degrade the efficiency of the exhaust aftertreatment system, resulting in an increase in output emissions. On the other hand, if too little air is pumped, excessive emissions may be generated because the percentage of fresh air to recirculated exhaust gas in the intake manifold becomes too high, resulting in an over-rich mixture that generates soot when combusted.

As the firing fraction calculator 70 transitions between different firing densities, the firing densities and associated fuel flows may change relatively quickly. However, due to the inherent lag of turbocharger system 90 in adjusting the intake manifold pressure to the new equilibrium level for the new firing density, the air flow per cylinder can only be adjusted gradually. Moreover, the ratio of recirculated exhaust gas to fresh air in the intake manifold can only be changed gradually due to the time required to adjust the EGR valve setting and the transit time of the exhaust gas through the system.

During the abrupt transition from the high firing density to the lower target firing density, one or more cylinders that are firing are skipped. This requires more fuel in each firing cylinder to maintain a constant torque level. Since the ratio of fresh air to recirculated exhaust gas in the intake manifold and the intake manifold pressure cannot be instantaneously changed, the air-fuel ratio will drop and become richer. For a fuel rich mixture, less torque may be produced due to incomplete combustion, and excessive soot may be produced due to insufficient fresh air to completely combust all of the injected fuel.

Controlling EGR system 52 to reduce the amount of recirculated exhaust gas in the intake manifold is one possible method of maintaining a more constant air-to-fuel ratio. By adjusting the mass of exhaust gas that is circulated back into the intake manifold, EGR system 52 may be used to achieve a desired ratio of fresh air to exhaust gas. The problem is that the adjustment does not take place immediately. Instead, multiple cylinder firing timings are typically required to achieve the desired ratio. During the temporary firing opportunity, the torque produced may be less than the torque required and the emissions may contain an excess of soot. Obviously, these are undesirable results and therefore require different control strategies.

Another possible approach is to transition from the first high firing density to the target lower firing density more slowly. For slower transitions in firing density, there may be sufficient time to purge excess exhaust gas from the intake manifold and allow the EGR system 52 to adjust the recirculated exhaust gas in the intake manifold to a desired ratio. However, relatively slow transitions in firing density may induce unacceptable levels of NVH. That is, the amount of vibration dose associated with the firing density transition may be unacceptable.

Similarly, an abrupt change in firing density from a low first firing density to a target high firing density will cause the previously skipped cylinder to fire. To maintain a constant torque, the fuel mass in each fired cylinder must decrease. Now, the air fuel has increased, becoming leaner, which may be undesirable. Depending on the initial air-fuel ratio, more NO may be generated during combustion_xOr the combustion may become unstable. Moreover, as more air passes through the engine 12, the temperature of the exhaust gas tends to decrease, which may degrade the efficiency of the exhaust aftertreatment system (e.g., 14A or 14B), thereby increasing unwanted emissions.

For spark ignition engines, the problems expressed explicitly above can be managed by a combination of: using a short firing density transition time, deactivating all skipped firing opportunities, and reducing torque output of the fired cylinder by retarding spark. As a result, no unburned air is pumped into the exhaust system, and the excess oxygen in the exhaust stream does not degrade aftertreatment. Spark control (e.g., late ignition) may be used to retard combustion timing, thereby reducing the torque output of the ignition and mitigating undesirable torque fluctuations (torque surge). Thus, a relatively smooth transition with acceptable NVH is achieved, as described in U.S. patent No. 9,745,905, which is incorporated herein for all purposes.

For compression ignition engines, there is no spark ignition and initiation of combustion is less controlled. However, compression ignition engines are not limited to operating at stoichiometric air-fuel ratios (as is often the case with spark ignition engines). Thus, more or less fuel may be injected to a point by simply injecting more or less fuel during the transitionTorque fluctuations or drops in firing density transitions are controlled in the firing cylinders. As described above, this may result in excessive soot or NO production in the exhaust_x. Also, in a compression ignition engine of stratified combustion (i.e., a diesel engine), noise, output torque, and exhaust gas temperature generated by combustion can be controlled via a fuel injection mode. For example, a diesel engine may use an injection pattern consisting of a pilot injection, a main injection, and a plurality of post injections. The pilot injection helps reduce combustion noise. The main injection produces the majority of the torque associated with cylinder firing. One or more post injections typically produce little torque, but may be used to increase the temperature of the exhaust. Thus, the fuel injection mode may be changed during the ignition density transition to help manage exhaust temperature, engine noise, output torque, and exhaust temperature. Due to a range of differences between spark ignition engines and compression ignition engines, the method for ignition density transition in spark ignition engines may not be optimal for compression ignition engines. In contrast, for existing compression ignition engines, managing the transition typically involves managing a compromise between acceptable NVH levels versus emission levels, and vice versa.

Managing firing density transitions

Applicants have found that by setting the firing density and pumping density of the cylinders separately during the transition, both excessive emissions and NVH can be avoided. In general, when operating at steady-state firing density, the firing density is equal to the pumping density, which implies that: in the event of misfire, no cylinder pumps air through the engine. For the present invention, skip fire controller 58 manages firing density transitions by separately defining (a) firing density and (b) pumping density. The pumping density should equal or exceed the firing density because each cylinder firing will cause a pumping event as well as a firing event. For a defined firing density, NVH may be reduced while the torque request is satisfied. For a defined pumping density, the rate of fresh air and/or recirculated exhaust gas pumped through the internal combustion engine is adjusted, allowing for better control of the ratio of fresh air to exhaust gas in the intake manifold. This may reduce emissions and improve fuel efficiency. Thus, by blending or interspersing cylinder firing with pumping of the unfueled cylinders, these conflicting requirements of meeting torque requirements, preventing excessive NVH, maintaining proper air/fuel ratio, and managing air flow through the engine may all be balanced, thereby producing more desirable results.

Referring to fig. 11, a flowchart 1100 illustrating steps for operating a lean burn internal combustion engine having separately defined ignition and pumping densities during a transition is illustrated.

In initial step 1102, the skip fire controller 58 operates the internal combustion engine at the current (i.e., starting or first) firing density as needed to meet the current torque demand.

In decision 1104, skip fire controller 58 determines whether the input parameters have changed sufficiently to warrant a change in firing density. As depicted in FIG. 3, the input parameters include torque request, air-to-fuel ratio, and signals from the aftertreatment element monitor. Additional input parameters not shown in FIG. 3 may include vehicle speed, engine speed, transmission gear ratio, oxygen sensor data, NO_xSensor data, ambient air temperature, exhaust temperature, barometric pressure, ambient humidity, and engine coolant temperature. If the input parameters have not changed sufficiently, skip fire controller 58 continues to operate the engine at the current firing density as provided in step 1102.

If the input parameters have changed sufficiently, skip fire controller 58 determines a new target firing and pumping density in step 1106. In general, when the firing density transition is complete, the pumping density will equal the firing density.

In steps 1108 and 1110, skip fire controller 58 updates the firing density (step 1108) and the pumping density (step 1110) for the next firing opportunity. The firing density and the pumping density may follow predetermined trajectories during the firing density transition. That is, for each firing opportunity, the values for firing density and pumping density may be determined from a look-up table, an algorithm, or by some other means.

In step 1112, the skip fire controller 58 defines the action for the next firing opportunity. That is, skip fire controller 58 determines whether the next firing opportunity will cause the cylinder associated with the firing opportunity to (a) fire, (b) pump without firing, or (c) be deactivated.

In step 1114, an ignition timing is executed. Depending on the outcome of step 1112, the cylinder associated with the firing opportunity is either (a) fired, (b) pumped but not fired, or (c) deactivated. Thus, for a typical 4-stroke engine, one of the following is performed:

(1) if ignited, then: (i) the method includes the steps of (i) introducing air from an intake manifold into the cylinder during an intake stroke, (ii) injecting fuel into the cylinder, (iii) compressing the air-fuel mixture during a compression stroke, (iv) combusting during a combustion stroke, and (v) expelling the combustion products and other gases out of the cylinder during an exhaust stroke.

(2) If pumped but not fired, the same sequence (i) to (v) is performed, but step (ii) is omitted. In the absence of fuel, no combustion occurs. As a result, during the exhaust stroke, the inducted air is pumped through the cylinder and into the exhaust manifold; and is

(3) If deactivated, no fuel is injected into the cylinder and either the intake valve(s) and/or the exhaust valve(s) are closed to prevent air from being pumped from the intake manifold through the engine into the exhaust manifold.

In step 1116, it is determined whether the transition to the target firing and pumping densities is complete. If not, process flow moves back to steps 1108 and 1110 where the firing density and pumping density are updated for the next firing opportunity. Steps 1112, 1114, and 1116 are then repeated for the next firing opportunity. On the other hand, if the transition to the target firing density and pumping density is complete, control returns to step 1102. The engine is operated at the target firing density until another change in torque demand or other input condition is determined that is sufficient to warrant the change in firing density. Possible other input conditions that may warrant a change in firing density include, but are not limited to, a change in aftertreatment element temperature, turbocharger setting, EGR setting, and air-fuel ratio.

It may be advantageous to operate a cylinder of an internal combustion engine as one of several different forms of gas springs during a duty cycle in which the cylinder is deactivated. This form of gas spring may include, but is not limited to, a Low Pressure Exhaust Spring (LPES), a High Pressure Exhaust Spring (HPES), or an Air Spring (AS). Different types of gas springs are formed by controlling the opening and closing sequence of the intake and exhaust valves during the deactivated work cycle. The low pressure exhaust spring is formed by exhausting exhaust from a previous firing duty cycle and then closing both the intake and exhaust valves in an immediately subsequent duty cycle. The high pressure exhaust spring is formed by trapping combusted exhaust gases in the cylinder during a deactivation duty cycle (by holding the exhaust valve closed at the end of the ignition duty cycle) following the ignition duty cycle. The air spring is formed by opening the intake valve during a deactivated duty cycle to introduce air into the cylinder but keeping the exhaust valve closed during the duty cycle and not firing the cylinder. These gas spring types are described in U.S. patent application No. 15/982,406, which is incorporated by reference herein for all purposes. It is also possible to deactivate a working cycle by closing only the intake or exhaust valves during the working cycle. In this case, gas is not trapped in the cylinder throughout the deactivated duty cycle, but may flow back and forth between the cylinder and the intake system (for a closed exhaust valve) or between the cylinder and the exhaust system (for a closed intake valve).

Referring to fig. 12A and 12B, a first example of a firing density transition from a high starting or first firing density to a lower target firing density is illustrated. In this example, the initial firing density is 1.0, and the firing density transitions to a target firing density of 0.5. This is the case where a momentary reduction in the number of active cylinders causes the air-fuel ratio (AFR) to be too rich due to excessive EGR.

Fig. 12A is a graph showing ignition density and pumping density during a transition. Sixty (60) ignition occasions are indicated along the horizontal axis. The vertical axis represents Firing Density (FD) or Pumping Density (PD). The transition of firing density and pumping density begins at the tenth firing opportunity, where both firing density and pumping density drop from their initial values of 1. The firing density and pumping density may follow a trajectory during the transition, which is illustrated in fig. 12A. Where the starting firing density is greater than the target firing density (e.g., 1.0>0.5), the transition is characterized such that (i) the firing density trajectory transitions to the target firing density in a first time period, and (ii) the pumping firing density trajectory transitions to its target pumping density in a second time period. As is apparent from fig. 12A, the first period of time is shorter than the second period of time. In this example, the firing density drops linearly until it reaches the target firing density of 0.5 at the firing opportunity 35. The firing density trajectory reaches a second firing density before the pumping density trajectory reaches its final value. The pumping density also decreases linearly, but the rate of change is more gradual than the firing density. The pumping density is not equal to the firing density until the firing opportunity 60, which in this example is the end of the transition. By transitioning firing density faster, NVH is reduced while continuing to meet torque demands. For slower pumping density changes, the rate of change of air flow through the internal combustion engine is slower, so adjustments to intake manifold pressure and mass of recirculated exhaust gas occur more slowly, better matching the natural time constants associated with these parameters. This causes the AFR to be closer to the desired AFR, thereby reducing emissions and providing more fuel efficient operation.

Fig. 12B plots the action associated with each firing opportunity along the same horizontal scale as in fig. 12A. Along the vertical axis, each of three possible firing opportunity actions or outcomes are indicated, including (1) fired, (2) pumped but not fired, and (3) deactivated.

In the particular example of FIG. 12B, initially all cylinders are fired and none are deactivated while operating at the first or starting firing density of 1.0. During the transition, individual cylinders are either fired according to firing density, pumped according to pumping density, or deactivated. For each of the sixty (60) firing occasions, the star symbol "+" indicates that the corresponding cylinder is (1) fired, (2) pumped, or (3) deactivated. In this example, there are seven firing occasions that result in pumping over the transition length of 50 firing occasions. The pumped firing opportunity occurs more often near the thirtieth to fortieth firing opportunities, where the difference between firing density and pumped density is greatest. At the completion of the transition (where the engine is operating at a target firing density of 0.5 and the pumping density is equal to the firing density), half of the firing opportunities are fired and the other half of the firing opportunities are deactivated. No firing opportunity results in pumping without firing.

Fig. 13A and 13B illustrate alternative examples in which the ignition density transition is increased during the transition. In this case, the transition is from an ignition density of 0.5 to 1.0. This is a transient change in firing density that causes the AFR to be too lean and thus increase NO_xExamples of situations of venting.

Fig. 13A plots firing density and pumping density versus firing timing during the transition. As in the previous example, the firing density and pumping density may follow a trajectory during the transition. In this example, at the start or first firing density, the target firing density is less than (e.g., 0.5)<1.0), the transition is characterized such that (i) the pumping density is initiated for the first time, and (ii) the change in firing density is retarded until after the change in pumping density is initiated. Due to the delay, the firing density trajectory remains at the first firing density, while the pumping density trajectory transitions toward its final value. However, once the firing density begins to change, it transitions faster than the pumping density. The firing density trajectory transition time is again shorter than the pumping density trajectory transition time, as is the transition from high firing density to low firing density. By minimizing the time at which the firing density transitions, the NVH associated with the transition may be minimized. By spreading the transition of pumping density, the airflow and exhaust gas recirculation ratio may better track the transition, thereby producing proper torque delivery without undue transitionNO of_xEmissions or NVH.

Fig. 13B plots the action associated with each ignition timing along the same horizontal scale as in fig. 13A. Initially, half of the firing occasions are fired and the other half of the firing occasions are deactivated. No firing opportunity results in pumping without firing. During sixty (60) firing occasions, each firing occasion may cause a cylinder to be fired, pumped, or deactivated. At the completion of the transition, the engine is operated at a target firing density of 1.0, which is also equal to the pumping density. All of the firing timings result in the act of firing the cylinders associated with the firing timing.

For each of the two examples described above, firing and deactivation of the cylinders is mixed or interspersed with pumping of the unfueled cylinders. As a result, these conflicting requirements of generating the required torque, minimizing NVH to acceptable levels, and controlling airflow for the purpose of reducing emissions can all be balanced, producing more desirable results for each.

Referring to FIG. 14, a graph of logic 140 for generating a fire enable flag and a pumping enable flag in response to a fire density signal and a pumping density signal during a transition is illustrated. The logic circuitry 140 includes a first sigma-delta converter 142, a second sigma-delta converter 144, and a logical or function 146. The first sigma-delta converter comprises an adder 141, an integrator or equivalently an accumulator 143, and a quantizer 145. The second sigma-delta converter comprises an adder 147, an integrator or equivalently an accumulator 149, and a quantizer 151.

For each firing opportunity during the transition, the first sigma-delta converter 142 receives the firing density signal value. The value may be an analog signal varying between 0 and 1. Examples of such signals are shown in fig. 12A and 13A. At each firing opportunity, the first sigma-delta converter 142 adds the received signal value to an accumulated value of previously received signal values stored in the accumulator 143 for the previous firing opportunity. Whenever the receipt of a new signal value causes the accumulated value to exceed the value of 1.0, the output quantizer 145 flips from zero to one (0 to 1), thereby setting the ignition enable flag. The output of the quantizer 145 is also fed back to the adder 141, which decrements the accumulated value by 1 whenever the ignition enable flag is set. In this way, the ignition enable flag is set whenever the accumulated value exceeds 1, and any remaining remainder is "rolled over" for the next ignition opportunity and saved in accumulator 143.

In a similar manner, the second sigma-delta converter 144 may receive a pumped density value in the form of an analog signal varying between 0 and 1. At each firing opportunity, the second sigma-delta converter 144 adds the received signal valve to an accumulated value of previously received signal values stored in the accumulator 149 for the previous firing opportunity. Whenever the receipt of a new signal value causes the accumulated value to exceed the value of 1, the output of the quantizer 151 rolls over from zero to one (0 to 1), thereby setting the intermediate pumping enable flag. In this way, the intermediate pumping enable flag 148 is set whenever the accumulated value exceeds 1.

The logical or function 146 is arranged to receive as inputs both the ignition enable flag and the intermediate pumping enable flag 148. For this arrangement, the OR function 146 generates a pumping enabled flag when either of these two inputs is true. The pump enable flag causes the cylinder associated with the firing opportunity to direct air during the work cycle. If both the pumping enabled flag and the ignition enabled flag are set, the ignition timing results in ignition of the cylinder. If only the pump enable flag is set, air is pumped through the cylinder without firing. If neither flag is set, the cylinder associated with the firing opportunity is deactivated, resulting in no air being directed or pumped. The output of the or function 146 is also fed back to an adder 147 that decrements the accumulated value by 1 whenever the ignition flag is set, and any remaining remainder is "rolled over" for subsequent signal values received for the next ignition opportunity.

For the fired cylinder, intake valve(s) are operated to allow air to be directed and fuel to be injected into the air to create a combustible air-fuel mixture, while exhaust valve(s) are operated to exhaust combustion gases. For pumping, intake and exhaust valves are operated to allow air to pass through the cylinder, but no fuel is provided. For deactivation, intake and/or exhaust valves are closed, thereby preventing pumping.

In one embodiment (as shown in fig. 14), the sigma-delta converters 142, 144 are first order sigma-delta converters. In other embodiments, the sigma-delta converters 142, 144 may be second, third, or higher order sigma-delta converters. Additionally, the logic 140 may be implemented in numerous ways, including but not limited to a microprocessor, a microcontroller, programmable logic, a field programmable gate array, discrete logic circuitry, hardware, firmware, or any combination thereof. Logic 140 may be implemented with different logic elements that are used in different arrangements but have the same or similar end results with respect to cylinder control.

Deceleration cylinder de-energizing (DCCO)

In certain driving situations, deceleration cylinder Deactivation (DCCO) occurs when there is no torque demand (e.g., the accelerator pedal is not depressed) by the driver or other autonomous or semi-autonomous driving controller, such as when the vehicle is coasting downhill or stopped. In DCCO, the cylinders of the engine are typically unfueled and the intake and/or exhaust valves are closed (i.e., deactivated). As a result, fuel is saved and pumping losses are reduced. Since cylinders in DCCO operation are typically not fired in numerous consecutive firing events, the temperature of the air in the intake manifold and/or the recirculated exhaust gas may increase. The temperature in the intake manifold may be measured or modeled. Either way, pumping through the engine may be used at DCCO exit to reduce the temperature of the gases in the intake manifold "on demand".

A real-world driving scenario where a DCCO exit strategy may be beneficial is mountain driving. When the vehicle is driven over long distance climbs, the torque demand on the engine will typically be very high for a long period of time, generating a large amount of waste heat and resulting in high engine operating temperatures. When the vehicle reaches a hill and begins to drive down the hill, the DCCO mode may be used if torque is not required, as would normally be the case when coasting downhill. DCCO operation may cause trapped gases in the intake manifold to become overheated because engine heat is absorbed due to a lack of pumping through the engine. Thus, when a torque demand is made (such as in the case of initiating another hill climb), the engine may not be able to deliver the required torque.

A possible solution to the above problem is to use the intake manifold gas temperature or an estimate thereof as a criterion for whether DCCO should be used or continued to be used. In the event that the intake manifold temperature is high, the DCCO is exited, or entry is disabled, and some cylinders pump air instead. That is, the pumping density is greater than zero when the firing density remains zero. As a result, cooler fresh air will flow into the intake manifold, thereby preventing heat build-up that would later limit torque production. Since pumping cooler air corresponds to when the engine and aftertreatment system are already hot, directing the cooler air should not significantly affect the efficiency of the aftertreatment system. When the intake manifold gas temperature is relatively low, then DCCO may be used as appropriate.

Another constraint on the use of DCCO in a diesel engine equipped with a turbocharger is to maintain a minimum turbocharger speed. If the turbocharger speed becomes too low or rotation stops, torque delivery will lag when torque is requested again. Furthermore, it is desirable to maintain a pressure differential across the turbocharger to avoid increased fuel consumption due to leakage from the compressor/turbine bearings. As a result, the duration of the DCCO event may be limited. Further, the engine may be operated by interspersing deactivated ignition timings with ignition timings for ignition and ignition timings that result in pumping air through the engine. This combination of actions associated with different firing timings may result in reduced fuel consumption because fewer cylinders are fired while maintaining sufficient turbine speed to deliver torque when torque is again requested.

Gasoline lean-burn engine

Much of the foregoing description has focused on stratified compression ignition engines (i.e., diesel engines); however, the present invention is not limited to diesel engines. In particular, gasoline-fueled engines may also be operated as lean-burn engines over a wide range of air/fuel ratios. Operation as a homogeneous-charge lean-burn engine can provide significant fuel efficiency gains when compared to stoichiometric operation. Homogeneous charge engines have an in-cylinder air-fuel mixture that is homogenously distributed throughout the in-cylinder volume. This air-fuel mixture may be ignited by either spark ignition or compression ignition.

FIG. 15 plots normalized specific fuel consumption on the vertical axis versus engine torque output on the horizontal axis for a gasoline fueled engine. Engine torque output is expressed in Net Mean Effective Pressure (NMEP). The data of the graph was collected with a representative boost 4-cylinder, 2.0-liter displacement engine operating at 2000rpm with spark ignition. Curve 302 depicts the fuel consumption for stoichiometric operation (λ 1.0) and curve 304 depicts the fuel consumption for lean operation (λ 1.7). Here, λ is an air-fuel ratio relative to a stoichiometric air-fuel ratio. Lambda values greater than 1 correspond to lean burn operation.

The graph of fig. 15 reveals that fuel efficiency can be improved by operating the engine with a homogeneous charge and an air-fuel ratio of λ 1.7 for engine torque outputs from approximately 3 to 14 bar. For engine outputs below 3 bar and above 14 bar, it is not feasible to operate at an ignition density of 1 and an air-fuel ratio of λ 1.7. In particular, for low engine output operation (<3 bar), an air-fuel ratio of λ 1.7 is not feasible due to combustion instability. For high engine output operation (>14 bar), an air-to-fuel ratio of λ 1.7 is not feasible due to insufficient airflow to provide the required torque output. In contrast, in these regions (i.e., <3 bar or >14 bar), stoichiometric operation (λ ═ 1.0) is preferred.

Fig. 15 was obtained with a fixed firing density operation of 1.0. By using an ignition density of less than 1, operation with a homogeneous lean charge can be extended to engine outputs below 3 bar, improving fuel efficiency in this range. In such an engine, there will be situations where a transition in air-to-fuel ratio is required to meet a changing engine torque demand. Such transitions may involve simultaneous changes in air-to-fuel ratio and ignition density. For example, to produce the same engine torque output, the engine may be operated at an ignition density of 0.5 and a stoichiometric air-fuel ratio (λ ═ 1.0), or at an ignition density of 1.0 and a lean air-fuel ratio of approximately 25:1(λ ═ 1.7).

The firing density may be selected in conjunction with air flow adjustments when a transition from stoichiometric to lean air/fuel ratios is desired. The air flow may be adjusted by changing intake manifold pressure, cam phase, valve lift, recirculated exhaust gas ratio, or some other means to achieve a desired air-to-fuel ratio from stoichiometry to a predetermined value. Intake manifold pressure may be regulated using a throttle and/or varying the level of boost provided by a turbocharger or supercharger. The fuel consumption may then be determined based on the torque request and the estimated energy conversion efficiency. By having the flexibility to adjust the firing density, a smooth transition of the in-cylinder air-fuel ratio from one value to another may be achieved without generating excessive spikes in exhaust emissions.

Fig. 16 shows the behavior of various engine parameters during an exemplary transition from an ignition density of 0.5 and a stoichiometric air-fuel ratio (λ ═ 1.0) to an ignition density of 1.0 and a lean air-fuel ratio of 25:1(λ ═ 1.7). In this example, it is assumed that the requested torque is substantially constant during the transition, and therefore the engine is controlled to deliver a substantially constant torque during the change in firing density.

Curve 200 illustrates the firing density transition from a firing density of 0.5 to a firing density of 1. The firing density transitions in a linear manner in the transition region; however, this is not a requirement; however, this is not a requirement.

Curve 202 illustrates the fuel flow for each fired cylinder during the transition. The fuel flow per firing cylinder drops by almost half because the firing event frequency doubles at the end of the transition as compared to the beginning of the transition. In contrast, the total engine fuel flow curve 204 is constant or substantially constant during the transition. In this case, it is slightly decreased because the engine is operated with a slightly higher efficiency at an ignition density of 1 compared to an ignition density of 0.5. This reflects the efficiency gain associated with the lean burn operation depicted in fig. 15.

Curve 206 illustrates the air flow per fired cylinder. The air flow per cylinder drops slightly during the transition.

In contrast, the total engine air flow (curve 208) approximately doubles during the transition. Since the ratio between the initial air-fuel ratio and the final air-fuel ratio is 1.7 in this case, the ratio between the initial air flow rate and the final air flow rate will also be approximated to this value. Adjustments may be made to the ratio of the total initial air flow to the final air flow based on different energy conversion efficiencies associated with different operating conditions.

Curve 210 illustrates a change in air-fuel ratio from its initial value of λ -1.0 to its final value of λ -1.7. The air-fuel ratio change between the first initial air-fuel ratio and the second final air-fuel ratio is gradually performed during the transition.

Curve 212 illustrates NO throughout the transition_xThe emissions remain at a low level.

Is incorporated by reference

U.S. patent No. 10,247,072 is incorporated herein by reference for all purposes.

Alternative embodiments

For certain embodiments of skip fire controller 58, the decision to fire or not fire a given cylinder of the engine is made dynamically, meaning on a firing opportunity by firing opportunity basis. In other words, a decision is made to either fire or skip a firing opportunity before firing each cylinder of the engine. Similarly, for certain embodiments of valve deactivation controller 102, the decision to allow or prevent pumping may be made dynamically on a firing opportunity by firing opportunity basis.

Intake and exhaust valve control may be more complex than simple binary control (i.e., opening or closing). The valve opening/closing timing may be adjusted using variable lift valves and/or may be adjusted by cam phasers. These actuators allow limited control of the cylinder MAC and its associated pumping losses without the use of a throttle. Advantageously, adjustment of the cylinder MAC allows control of the air-fuel ratio for a fixed fuel charge. The combustion conditions may then be optimized to improve fuel efficiency or to provide other desired conditions in the combustion exhaust, i.e., temperature, pollutant emission levels, etc.

The invention has been described primarily in the context of controlling ignition in a 4-stroke compression-ignition piston engine suitable for use in a motor vehicle. Compression ignition may be in a stratified fuel charge, a homogeneous fuel charge, a portion of a homogeneous mass, or some other type of fuel charge. However, it should be appreciated that the described skip fire approach is well suited for use in a wide variety of internal combustion engines, including gasoline and/or Spark Ignition (SI) type engines. Additionally, any of the engines described herein may be used in virtually any type of vehicle-including automobiles, trucks, locomotives, ships, boats, construction equipment, airplanes, motorcycles, scooters, etc.; and virtually any other application involving cylinder ignition and utilizing internal combustion engines.

Although only a few embodiments have been described in detail, it should be understood that the present application may be embodied in many other forms without departing from the spirit or scope of the disclosure provided herein. For example, as described above, two sigma-delta converters are used to determine the action associated with the firing opportunity in the firing density transition. This is not a requirement. In one non-exclusive embodiment, skip fire controller 58 may rely on one or more look-up tables to obtain the action associated with any given firing opportunity. For a given combination of starting firing density and target firing density, the look-up table may provide possible firing patterns and pumping patterns associated with each firing density transition. The invention described herein may be applied when both the starting firing density and the target firing density in the firing density transition are non-zero, and may also be applied when one of the starting firing density or the target firing density is zero (i.e., all cylinders are skipped). Accordingly, the present embodiments are to be considered as illustrative and not restrictive, and the details given herein are not to be limited, but may be modified within the scope and equivalents of the appended claims.

Claims (83)

1. An engine controller in a vehicle for controlling a lean-burn internal combustion engine, the controller configured to:

while operating the internal combustion engine at a first firing density, determining a target firing density suitable for meeting a requested torque demand for the internal combustion engine;

determining an ignition density during a transition from the first ignition density to the target point ignition density;

determining a pumping density during a transition from the first firing density to the target point firing density;

operating the internal combustion engine during the transition by:

selecting to ignite or not to ignite each ignition opportunity according to the determined ignition density;

selectively (a) pumping or (b) not pumping air through the internal combustion engine based on the determined pumping density for each firing opportunity that is not fired.

2. An engine controller as recited in claim 1 wherein firing density and pumping density during the transition are determined to meet requested torque demand while balancing operating requirements at acceptable noise, vibration and harshness (NVH) levels and minimizing exhaust emissions.

3. An engine controller as claimed in claim 1 or 2, further configured to adjust individually:

the firing density to mitigate excessive NVH; and

the pumping density to mitigate excessive emissions caused by too much or too little air flow through the internal combustion engine.

4. An engine controller according to any one of claims 1 to 3, when the first firing density is greater than the target firing density, the engine controller is further configured to:

transitioning the firing density from the first firing density to the target firing density in less time than transitioning the pumping density, wherein:

(a) this faster transition in firing density reduces NVH; and is

(b) The slower transition in the pumping density reduces the rate of change of the air flow through the internal combustion engine.

5. An engine controller according to any one of claims 1 to 4, when the first firing density is less than the target firing density, the engine controller is further configured to:

(a) initiating a transition in the pumping density; and

(b) the transition in firing density is retarded until after the transition in pumping density has begun.

6. An engine controller according to any one of claims 1 to 5, wherein the fuel injection pattern is changed during a transition from the first firing density to the target point firing density.

7. An engine controller as recited in any of claims 1-6 wherein the determined pumping density is used to at least partially control the mass of air pumped from the air intake manifold through the internal combustion engine and/or recirculated exhaust gas.

8. An engine controller as recited in any of claims 1-7 wherein the mass of air and/or recirculated exhaust gas pumped through the internal combustion engine is proportional to the pumping density.

9. An engine controller as recited in any of claims 1-8 further comprising a first sigma-delta converter arranged to generate a spark enable flag for each cylinder firing command as a function of the determined firing density.

10. The engine controller of claim 9, further comprising a second sigma-delta converter arranged to generate an intermediate pumping enabled flag for each cylinder pumping command as a function of the determined pumping density.

11. An engine controller as recited in claim 10 further comprising a logic or function for performing a logic or operation between the ignition enable flag and the intermediate pumping enable flag to generate a pumping enable flag when either or both of the intermediate pumping enable flag or the ignition enable flag is a logic true.

12. An engine controller according to claim 11, wherein the logical or function is further configured to: when both the intermediate pumping enabled flag and the spark enabled flag are logically false, causing a spark timing results in deactivating the cylinder associated with the spark timing.

13. An engine controller as recited in any of claims 1-12 further configured to: upon exiting a deceleration cylinder Deactivation (DCCO) event, the pumping density is increased to induce more pumping of air through the internal combustion engine.

14. An engine controller as recited in any of claims 1-13 wherein the determined pumping density is greater than the determined firing density for some firing occasions in the transition.

15. An engine controller according to any one of claims 1 to 14, further comprising: the firing density is determined on a firing opportunity by firing opportunity basis during a transition from the first firing density to the target point firing density.

16. An engine controller according to any one of claims 1 to 15, further comprising: the pumping density is determined on a firing opportunity by firing opportunity basis during a transition from the first firing density to the target point firing density.

17. An engine controller arranged to operate a lean burn internal combustion engine in a skip fire mode, the engine controller being arranged to define a firing density profile and to define a pumping density profile when transitioning from a first firing density to a second firing density different from the first firing density, the defined firing density profile (a) meeting a torque request and (b) producing acceptable noise, vibration and harshness (NVH) levels, and the defined pumping density profile controlling an air pumping rate through the internal combustion engine.

18. An engine controller as recited in claim 17 further configured such that when the first firing density is greater than the target firing density, the transition is characterized by the firing density trace reaching the second firing density before the pumping density trace reaches its final value.

19. An engine controller according to claim 17 or 18, further configured such that when the first firing density is less than the target firing density, the transition is characterized by the firing density trace remaining at the first firing density and the pumping density trace transitioning towards its final value.

20. An engine controller as recited in any of claims 17-19 further comprising a first sigma-delta converter for generating a spark enable flag as a function of the firing density trajectory.

21. The engine controller of claim 20, further comprising a second sigma-delta converter for generating an intermediate pumping enabled flag according to the pumping density profile.

22. An engine controller as recited in claim 21 further comprising an or function for generating a pumping activation flag that receives (a) the ignition activation flag and (b) the intermediate pumping activation flag and generates the pumping flag when either (a) or (b) is true.

23. An engine controller according to claim 22, further configured to deactivate the cylinders during an ignition timing when the pumping activation flag is not set.

24. An engine controller as recited in any of claims 17-23 further configured such that the firing density trajectory transition time is shorter than the pumping density trajectory transition time.

25. An engine controller arranged to operate a lean-burn internal combustion engine, the engine controller being configured to:

receiving an input signal indicative of a temperature of air within an intake manifold arranged to provide air to the internal combustion engine;

(a) operating the lean-burn internal combustion engine in a deceleration cylinder Deactivation (DCCO) mode when both of: (i) the input signal indicates that the air temperature in the intake manifold is below a threshold temperature, and (ii) the internal combustion engine has not received a torque request; and is

(b) Not operating the lean burn internal combustion engine in the DCCO mode when (iii) the input signal indicates that the temperature of air within the intake manifold is above the threshold temperature or (iv) the internal combustion engine receives a torque request.

26. An engine controller according to claim 25, further configured to: inducing pumping of air through the lean-burn internal combustion engine when the internal combustion engine does not receive a torque request but the temperature of air within the intake manifold is above the threshold temperature.

27. An engine controller according to claim 25 or 26, wherein the input signal is generated in response to a measure of the temperature of air in the intake manifold.

28. An engine controller according to any one of claims 25 to 27 wherein the input signal is generated in response to the output of a model that estimates the temperature of air within the intake manifold.

29. A method of transitioning firing density from a first firing density to a target second firing density different from the first firing density in a skip fire controlled internal combustion engine in response to a change in an input parameter, the method comprising:

causing at least one first ignition timing to be an ignition timing of ignition during the transition;

causing at least one second ignition timing to be a pumped ignition timing during the transition, wherein air is pumped through the engine without combusting the air; and

at least one third ignition timing is caused to be a deactivated ignition timing during the transition, wherein no air is pumped through the engine.

30. The method of claim 29, wherein the input parameter is selected from the group consisting of: torque request, air-to-fuel ratio request, aftertreatment element signal, vehicle speed, engine speed, transmission gear ratio, oxygen sensor signal, NO_xA sensor signal, an ambient air temperature signal, an exhaust temperature signal, a catalyst temperature signal, an atmospheric pressure signal, an ambient humidity signal, and an engine coolant temperature signal.

31. A method according to claim 29 or 30, wherein during the transition:

(i) the pumping density is greater than the firing density; or

(ii) The pumping density and the firing density are the same.

32. A method according to any one of claims 29 to 31 wherein the internal combustion engine is a lean burn engine.

33. A method according to any one of claims 29 to 32 wherein the internal combustion engine is a compression ignition engine.

34. A method according to any one of claims 29 to 33 wherein the firing occasions of the firings, pumping firing occasions and deactivating firing occasions are interspersed in the transition.

35. The method of any of claims 29-33, wherein the pattern of firing occasions of firing, pumping and deactivating during the transition is determined at least in part by a sigma-delta converter.

36. A method according to any of claims 29 to 36, wherein the pattern of firing occasions on fire, pumping firing occasions and deactivating firing occasions during the transition is determined at least in part by a look-up table.

37. A method of making an ignition density transition in a skip fire controlled gasoline fueled internal combustion engine, the method comprising:

operating at a first firing density and an associated first air-to-fuel ratio; and

initiating a transition to a second firing density that is different from the first firing density and has an associated second air-to-fuel ratio;

wherein the second air-fuel ratio is different from the first air-fuel ratio.

38. The method of claim 37, further comprising: the air-fuel ratio is gradually changed between the first air-fuel ratio and the second air-fuel ratio during the transition.

39. The method of claim 37 or 38, wherein one of the first air-fuel ratio or the second air-fuel ratio is a stoichiometric air-fuel ratio.

40. A method according to claim 36 or 37 wherein one of the first air-fuel ratio or the second air-fuel ratio is a lean air-fuel ratio.

41. The method of any of claims 37 to 40, further comprising: a homogeneous charge is formed within a working chamber of the internal combustion engine when operating at a lean air-fuel ratio, wherein the lean air-fuel ratio is one of the first air-fuel ratio or the second air-fuel ratio.

42. The method of any of claims 37 to 41, further comprising: operating the internal combustion engine with spark ignition at the first ignition density or the second ignition density.

43. The method of any of claims 37 to 42, further comprising: operating the internal combustion engine with spark ignition at both the first ignition density and the second ignition density.

44. The method of any of claims 37 to 43, further comprising: operating the internal combustion engine at the first firing density with compression ignition and at the second firing density with spark ignition.

45. An internal combustion engine having a plurality of cylinders, each of the cylinders being arranged to combust an air-fuel mixture characterized by an air-fuel ratio, the internal combustion engine comprising:

a skip fire controller arranged to: operating the internal combustion engine in a skip fire mode by determining a firing density to selectively operate the internal combustion engine at a reduced effective displacement of the internal combustion engine that is less than a maximum displacement,

the skip fire controller determines the firing density using a combination of:

(a) a temperature of exhaust gas discharged from the internal combustion engine;

(b) an air-fuel ratio of the air-fuel mixture contained in one or more of the cylinders; and

(c) the torque is requested in a torque-requesting manner,

wherein at least one cylinder of the internal combustion engine is fired, skipped, and selectively fired or skipped in successive firing occasions when operating at the determined firing density.

46. The internal combustion engine according to claim 45, wherein the engine controller is further configured to dynamically determine different firing densities as (a), (b), and/or (c) change in response to changing driving conditions, wherein each of the different firing densities indicates a different reduced effective displacement, all of the reduced effective displacements being less than a maximum displacement of the internal combustion engine.

47. The internal combustion engine of claim 45 or 46, wherein the engine controller is further configured to: adjusting a temperature of exhaust gas passing through an aftertreatment system when the internal combustion engine is operating in the skip fire mode.

48. An internal combustion engine according to any one of claims 45 to 47, wherein the engine controller adjusts the temperature of exhaust gas passing through the aftertreatment system in the skip fire mode by pumping air through at least one skipped cylinder, the pumped air being used to reduce the temperature of exhaust gas passing through the aftertreatment system.

49. The internal combustion engine of claim 48, wherein during skipped firing opportunities, air is pumped through the skipped cylinder by opening intake and exhaust valves, but the skipped cylinder is not fueled.

50. An internal combustion engine according to any one of claims 45 to 49, wherein the engine controller adjusts the temperature of exhaust gas passing through the aftertreatment system in the skip fire mode by preventing pumping air through at least one skipped cylinder, the lack of pumped air being used to prevent a decrease in temperature of exhaust gas passing through the aftertreatment system.

51. The internal combustion engine of claim 50, wherein pumping air through the at least one skipped cylinder is prevented by:

(d) closing the intake valve;

(e) closing the exhaust valve; or

(f) Both (d) and (e).

52. The internal combustion engine according to any one of claims 45 to 51, wherein the engine controller is further configured to adjust the exhaust gases to within a temperature range.

53. The internal combustion engine according to claim 52, wherein the temperature range is from substantially 200 ℃ to substantially 400 ℃.

54. The internal combustion engine according to claim 52, wherein the engine controller is further configured to adjust the exhaust gases to within the temperature range by:

(a) allowing or increasing pumping air through the internal combustion engine when the exhaust gases are above a threshold temperature within the temperature range; and is

(b) Pumping air through the internal combustion engine is prevented or reduced when the exhaust gases are below a threshold temperature within the temperature range.

55. The internal combustion engine according to claim 52, wherein the aftertreatment system is arranged to reduce nitrogen oxides (N)O_x) And (5) discharging.

56. The internal combustion engine of any one of claims 45 to 55, further comprising an Exhaust Gas Recirculation (EGR) controller arranged to at least partially define an air-to-fuel ratio of an air-fuel mixture contained within and combusted by the plurality of cylinders, the EGR arranged to determine the air-to-fuel ratio based at least in part on an ignition density determined by the engine controller.

57. An internal combustion engine according to any one of claims 45 to 56, further comprising a boost controller arranged to define at least in part an intake pressure of the plurality of cylinders, the boost controller being arranged to define the intake pressure based at least in part on an ignition density determined by the engine controller.

58. An internal combustion engine according to any one of claims 45 to 57, wherein the engine controller is further configured to operate the internal combustion engine in the skip fire mode by:

(d) operating the internal combustion engine at a relatively high firing density;

(e) targeting operation of the internal combustion engine at a relatively low firing density in response to a change in the torque request; and

(f) gradually transitioning operation of the internal combustion engine from the relatively high ignition density to the relatively low ignition density using one or more intermediate ignition densities, the gradual transition mitigating a surge in particulate emissions of the internal combustion engine.

59. An internal combustion engine according to any one of claims 45 to 58, wherein the engine controller is further configured to operate the internal combustion engine in the skip fire mode by:

(d) operating the internal combustion engine at a relatively low firing density;

(e) targeting operation of the internal combustion engine at a relatively high firing density in response to a change in requested torque demand; and

(f) gradually transitioning operation of the internal combustion engine from the relatively low ignition density to the relatively high ignition density using one or more intermediate ignition densities, the gradual transition mitigating NO of the internal combustion engine_xThe emission is dramatically increased.

60. The internal combustion engine according to any one of claims 45 to 59, further comprising a valve deactivation control unit for selectively opening or closing intake and/or exhaust valves, respectively, associated with selected ones of the plurality of cylinders of the internal combustion engine, wherein the valve deactivation control unit is for selectively opening or closing intake and/or exhaust valves of the selected cylinders during skipped firing opportunities based on whether it is desired to pump air into the exhaust system.

61. The internal combustion engine of claim 60, wherein the valve deactivation control unit makes the decision to selectively open or close the intake and/or exhaust valves associated with the selected cylinder based at least in part on the intake pressure generated by the boost system.

62. The internal combustion engine according to claim 60, wherein the valve deactivation control unit makes the decision to selectively open or close the intake and/or exhaust valves associated with the selected cylinder based at least in part on the actual or estimated temperature of the exhaust gases.

63. An internal combustion engine according to any one of claims 45 to 59, further comprising a valve deactivation control unit arranged to selectively open or close intake and/or exhaust valves associated with the plurality of cylinders of the internal combustion engine, respectively, the selective opening or closing being based on a prioritization scheme defining a priority order comprising:

(1) the turbocharger operates at or near surge conditions; and

(2) the actual or estimated temperature of these exhaust gases is outside of the normal operating range; and

(3) the current torque request.

64. The internal combustion engine of claim 63, wherein the valve deactivation control unit is further arranged to prioritize selective opening of intake and exhaust valves associated with one or more of the plurality of cylinders based on whether the compressor is operating at or near the surge condition.

65. The internal combustion engine according to claim 63 or 64, wherein the valve deactivation control unit is further arranged to prioritize the selective opening or closing of intake and/or exhaust valves associated with one or more of the plurality of cylinders based on whether the actual or estimated temperature of the exhaust gases is too high or too high relative to the normal temperature range defined as being suitable for reducing NO_xDischarging the waste water, and discharging the waste water,

wherein intake and exhaust valves associated with the one or more cylinders are opened to induce pumping when the temperature is too high;

wherein intake and/or exhaust valves associated with the one or more cylinders are closed to prevent pumping when the temperature is too low.

66. An internal combustion engine according to any one of claims 45 to 65, wherein the internal combustion engine is one of:

(d) a compression ignition engine;

(e) a spark ignition engine;

(f) a lean burn engine;

(g) a diesel engine; or

(h) A gasoline engine.

67. An internal combustion engine according to any one of claims 45 to 66, further comprising an aftertreatment system for treating exhaust gas emitted from the internal combustion engine, the aftertreatment system comprising:

(e) a particulate filter;

(f) an oxidation catalytic converter;

(g) a reduction catalyst converter; or

(h) Any combination of (e) to (g).

68. An internal combustion engine according to any one of claims 45 to 67, wherein the skip fire controller is a dynamic skip fire controller arranged to make decisions to fire or skip the at least one cylinder dynamically on a firing opportunity by firing opportunity basis.

69. The internal combustion engine of any one of claims 45 to 68, wherein the internal combustion engine powers a vehicle.

70. An internal combustion engine according to any one of claims 45 to 69, wherein the air-fuel ratio in one of the plurality of cylinders is adjusted based on the ignition history of the cylinder.

71. The internal combustion engine of claim 70, wherein the timing of fuel injection or the pattern of fuel injection is adjusted based on the ignition history of the cylinder.

72. An internal combustion engine according to any one of claims 45 to 71, wherein the air-fuel ratio in one of the plurality of cylinders is adjusted based on an engine skip fire sequence prior to firing the cylinder.

73. The internal combustion engine of claim 72, wherein timing of fuel injection or a pattern of fuel injection is adjusted based on the engine skip fire sequence prior to firing the cylinder.

74. A method of performing an ignition density transition in a boosted lean burn internal combustion engine with skip fire control, the method comprising:

operating at a first firing density; and

initiating a transition to a second firing density, the second firing density being different from the first firing density;

wherein the air-fuel ratio is maintained within a predefined range during the transition.

75. The method according to claim 74 wherein the predefined air-to-fuel ratio is between 20 and 35.

76. A method according to claim 74 or 75 wherein the air-fuel ratio is maintained within a predefined range by adjusting an engine parameter selected from the group consisting of: ignition density, boost level, exhaust gas recirculation level, intake valve lift profile, and exhaust valve lift profile.

77. A method of controlling output torque in a lean burn internal combustion engine having skip fire control of a plurality of cylinders, the method comprising:

receiving a request for the output torque;

controlling an amount of fuel delivered to each cylinder in the internal combustion engine to cause the internal combustion engine to deliver a requested output torque;

the mass of air inducted into each cylinder is controlled to cause the air-fuel ratio within that cylinder to fall within a predefined range.

78. A method as claimed in claim 77 wherein controlling the quality of the directed air is achieved by adjusting an engine parameter selected from the group consisting of: ignition density, boost level, exhaust gas recirculation level, intake valve lift profile, and exhaust valve lift profile.

79. A method according to claim 77 or 78, wherein the firing density is selected from a plurality of firing density levels.

80. A method according to any one of claims 77 to 79, wherein the internal combustion engine is operated at an intermediate firing density level during the transition from the first firing density level to the second, different, firing density level.

81. A method as claimed in any one of claims 77 to 80 wherein during the transition from the first firing density level to the second different firing density level, the transition is spread over a plurality of engine cycles.

82. A method according to any one of claims 77-81, wherein the firing density level is adjusted to maintain the temperature of exhaust gas from the internal combustion engine within a predefined range.

83. The method according to any one of claims 82, wherein the predefined air-to-fuel ratio is between 20 and 35.

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