EP1418330A1

EP1418330A1 - A method of controlling an internal combustion engine

Info

Publication number: EP1418330A1
Application number: EP20020025180
Authority: EP
Inventors: Rob Otterspeer; Maria Segerling; Carina Björnsson
Original assignee: Ford Global Technologies LLC
Current assignee: Ford Global Technologies LLC
Priority date: 2002-11-11
Filing date: 2002-11-11
Publication date: 2004-05-12
Anticipated expiration: 2022-11-11
Also published as: EP1418330B1; DE60209614T2; DE60209614D1

Abstract

A method in an internal combustion engine is presented. Controllable valves and controllable air and fuel input means are provided at each cylinder. The method includes controlling the valves and the air and fuel input means so that a transition is made from a first stroke mode to a second stroke mode. In the first and second stroke mode the crankshaft rotates a first and a second angular distance, respectively, between two consecutive ignitions of the engine. At least one transition combustion load at or near the transition is controlled, the transition combustion load being determined at least partly on the basis of the first and/or the second angular distance.

Description

TECHNICAL FIELD

The present invention relates to a method in an internal combustion engine including a crankshaft, whereby controllable valves and controllable air and fuel input means are provided at each cylinder. The method comprises the steps of controlling the valves and the air and fuel input means so that the engine is operating in a first stroke mode, in which the crankshaft rotates essentially a first angular distance between two consecutive ignitions of the engine, and controlling the valves and the air and fuel input means so that the engine is operating in a second stroke mode, in which the crankshaft rotates essentially a second angular distance between two consecutive ignitions of the engine, the second angular distance being non-equal to the first angular distance.

BACKGROUND

In internal combustion engines it is possible to achieve operation in different stroke modes, i.e. the cylinders can from time to time be operated in different stroke cycles. For example, the engine can be operated in a four stroke mode or an eight stroke mode, by use of cylinder deactivation. However, in engines with valves dependent on the crankshaft motion, only stroke modes being multiples of four, i.e. four, eight, twelve, sixteen, etc. can be obtained. As opposed to this, in engines where the valves are controlled freely, i.e. independent of the crankshaft, other stroke modes, including six stroke, can be achieved.
In some transitions from one stroke mode to another, problems occur during the transition period. These problems are partly due to the fact that possible ignition events are restricted to instances where the pistons are in their top dead center (TDC), and the positions of the pistons in relation to the crankshaft are fixed. Thus, in these cases special consideration have to be taken to achieve a transition in which the operation of the engine is not adversely affected and, in the case of vehicle propulsion, driver and passengers are not subjected to any discomfort.

SUMMARY OF THE INVENTION

It is an object of the present invention to make it possible to change the stroke modes of an internal combustion engine, at which the operation of the engine is undisturbed.
It is another object of the present invention to make it possible to change the stroke modes of an internal combustion engine in a vehicle, without compromising the drivability of the vehicle.
It is also an object of the present invention to improve the comfort of the driver and passengers in a vehicle powered by an internal combustion engine, when the engine changes the stroke modes.
These objects are reached by a method of the type described initially, further comprising controlling the valves and the air and fuel input means so that a transition is made from the first stroke mode to the second stroke mode, and controlling at least one transition combustion load at or near the transition, the transition combustion load being determined at least partly on the basis of the first and/or the second angular distance.
Thereby, any deviations of the engine output torque due to irregular engine ignition intervals in the transition can be attended to by adjusting combustion loads to counteract the influence of changing intervals.
Preferably, in the transition, where the crankshaft rotates essentially a third angular distance between two consecutive ignitions of the engine, the third angular distance being non-equal to the first and the second angular distances, the transition combustion load is determined at least partly on the basis of the third angular distance in relation to the first angular distance and/or the second angular distance.
The situation where three ignition interval distances occur during and around a transition appears in changes from four to six stroke modes in engines with an odd number of cylinders. By determining a combustion load based on one of the distances in relation to at least one of the two others, disturbances in the engine output torque due to the non-equalness of the ignition interval distances at the transition can be counteracted. In, turn this will improve the performance of the engine, and if the latter is used for vehicle propulsion, secure the comfort to the driver and passengers, as well as support the drivability of the vehicle.
Preferably, the transition combustion load is related to an ignition occurring at the beginning or the end of the third angular distance rotation of the crankshaft. Thereby, the transition combustion load is located close in time to the time interval in which the crankshaft rotates the third angular distance. As a result a minimum time elapses between the cause of the problem, i.e. an "off-mode" ignition interval, and its remedy according to the invention, i.e. a combustion load adapted to compensate for the dissimilar ignition interval.
Preferably, where one of the transition combustion loads is related to an ignition at which the crankshaft rotates essentially the first angular distance between the ignition and an immediate preceding ignition, and the crankshaft rotates essentially the second angular distance between the ignition and an immediate following ignition, the transition combustion load is determined at least partly on the basis of the first and the second angular distance.
In cases of one ignition interval distance being followed by another, combustion loads adapted to any of the stroke modes, being applied for all ignition in the transition, will provide a momentary drop or increase in the engine output torque, as described in detail below. By adjusting the combustion load at the ignition between the two modes, the drop or increase of the torque can be counteracted. This provides for a smooth operation of the engine in transition, also in cases where only two ignition interval distances occur in the transition.

DESCRIPTION OF FIGURES

Below, the invention will be described in detail with the aid of the drawings, in which

fig. 1 shows schematically a part of a longitudinal cross-section of a cylinder in an internal combustion engine, and a control device depicted as a block,
fig. 2 and 3 show diagrams of the piston movement and ignition in each cylinder in a five cylinder engine, as a function of the crankshaft angle, in four and six stroke mode, respectively,
fig. 4 shows a diagram of the piston movement and ignition in each cylinder in a five cylinder engine, as a function of the crankshaft angle, at a transition from four stroke mode to six stroke mode,
fig. 5 shows a diagram of the piston movement and ignition in each cylinder in a five cylinder engine, as a function of the crankshaft angle, in eight stroke mode, and
fig. 6 shows a diagram similar to the one in fig. 4, for a transition from six stroke mode to eight stroke mode.

DETAILED DESCRIPTION

For this presentation, the expression "stroke cycle" refers to the working cycle in each cylinder of an engine. For example, a four stroke cycle includes the strokes compression (ignition), expansion, expulsion and induction. The expression "stroke mode" refers to the operational mode of the engine regarding the stroke cycle in which the cylinders work. For example, if the cylinders work in a four stroke cycle, then the engine is in four stroke mode. The expression "engine ignition interval" refers to the interval between two consecutive ignitions in the engine. In normal operation, these consecutive ignitions each occur in different cylinders of the engine.
Fig. 1 shows schematically an arrangement at a cylinder 11 of an inline five cylinder internal combustion engine. A similar arrangement is shown in PCT/SE99/01947, incorporated herein by reference.
A piston 12 is connected to a crankshaft (not shown) via a piston rod 13. At each cylinder two, three, four, five or more valves are provided. In fig 1 only two valves 14, 15 are shown, an exhaust valve 14 and an intake valve 15. The movement of each valve can be controlled individually by a control device 16. Activation of the valves can be done with hydraulic, pneumatic, electromagnetic, piezoelectric or any other known activation aid.
Thus, the valves can be moved independently of the crankshaft. As explained further below, the engine can be run in different stroke modes, and the possibility to control the valves independently of the crankshaft is advantageous for facilitating different stroke modes of the engine. However, certain types of stroke modes, e.g. four and eight stroke mode, can be achieved with traditionally camshaft activated valves. In the latter case well known cylinder deactivation techniques are being used.
Air and fuel input means comprising a fuel injector 17 are provided at each cylinder and controllable by the control device 16. By controlling the fuel input means the combustion load at each ignition can be controlled. The fuel injector 17 can be located in the intake part at the cylinder, as depicted in fig. 1, or in the combustion chamber of the cylinder. The intake valve 15 could be part of the air and fuel input means and used to control the amount of gas inducted. Alternatively, the air and fuel input means could comprise a butterfly valve in the intake to the cylinder. Thereby, throttling can be performed with the butterfly valve, or with a combined use of the butterfly valve and the intake valve 15.
It should also be noted that the fuel and air input means can include known arrangements for supercharging the medium injected into the cylinder, e.g. turbo charge or compressor, etc.
Fig. 1 also depicts igniting means 18 in each cylinder, comprising a spark plug.
During engine operation, the valves and the air and fuel input means can be controlled so that the engine is operating in a first stroke mode. In the first stroke mode, the valves and the air and fuel input means at each cylinder are controlled so that a first stroke cycle is performed at each cylinder.
As an example, the first stroke cycle is a four stroke cycle, with the strokes compression (and ignition), expansion, expulsion, and induction. Denoting the cylinders of the engine 1,2,3,4,5 with respect to their relative spatial position, a suitable firing order between the cylinders is 1,2,4,5,3. However, any alternative firing order can be used.
Fig. 2 shows a diagram of the location of the piston, (sine curves), and ignitions, (large dots), in each cylinder in the five cylinder engine, as a function of the crankshaft angle, in a four stroke mode, which is the first stroke mode in this example. The ignitions are indicated with large dots, and it can be seen that the interval between each ignition of the engine is 144 degrees of crankshaft angle, which is the length of the four stroke cycle, 720°, divided by the number of cylinders: five.
During engine operation, the valves and the air and fuel input means can be controlled so that the engine is operating in a second stroke mode. In the second stroke mode, the valves and the air and fuel input means at each cylinder are controlled so that a second stroke cycle is performed at each cylinder.
In the example here, the second stroke cycle is a six stroke cycle. Thereby, the stroke order of each cylinder could be: compression (and ignition), expansion, expulsion, induction, compression, and expansion. This stroke order is suitable for conditions under normal operating temperature of the engine. The additional compression of inducted gases before final compression and ignition increases the duration of mixing of fuel and air with 200 percent, which in turn enhances combustion performance.
Alternatively, the stroke order of each cylinder in a six stroke cycle could be: compression (and ignition), expansion, compression, expansion, expulsion, and induction. Such a stroke order is suitable in cold start operation, since the repeated compression and expansion after ignition increases the heat transportation to the cylinder walls and accelerates heating of the engine.
Regardless of the stroke order used, if the engine is operating in a low stroke mode and at a relatively low external load, the low combustion load at each ignition will result in a low efficiency, largely due to a low combustion temperature. At the same external load of the engine, a higher stroke mode will result in a higher efficiency at each combustion. This is because larger combustion loads are used combined with longer ignition intervals resulting in higher combustion temperatures.
In six stroke mode, a suitable firing order between the cylinders is 1,3,5,4,2. However, any alternative firing order can be used. Fig. 3 shows a diagram of the location of the piston, (sine curves), and ignitions, (large dots), in each cylinder, as a function of the crankshaft angle, in the six stroke mode, which is the first stroke mode in this example. It can be seen that the interval between each ignition of the engine is 216 degrees of crankshaft angle, which is the length of the six stroke cycle, 1080°, divided by the number of cylinders: five.
Thus, in a five cylinder engine, and also in other engines with odd number of cylinders, it is possible to achieve equal distances between ignitions, in both four stroke mode and six stroke mode. In addition, eight stroke mode and higher stroke modes with equidistant ignitions can be obtained with odd cylinder engines, see below. In even cylinder engines, eight stroke mode and twelve stroke mode with equidistant ignitions are achievable. However, six stroke mode and ten stroke mode with equidistant ignitions are not achievable in engines with an even number of cylinders.
According to the invention, the valves 14, 15 and the air and fuel input means 17 can be controlled so that a transition is made from the first stroke mode to the second stroke mode and so that, during the transition, a compressed mixture of air and fuel is present in each cylinder at each ignition.
That the mixture of air and fuel is compressed, implies that each ignition takes place near a Top Dead Center (TDC) of the respective cylinder. It also implies that the respective valves are closed at each ignition. Since it takes four strokes to expel the exhaust, to induct a new fresh mixture and to compress the latter, a preceding ignition in a certain cylinder has to have taken place at least four strokes before an ignition in the transition, (and outside the transition as well).
For an understanding of the procedure at a mode transition, we refer to fig. 4, which shows a diagram of the location of the piston, (sine curves), and ignitions, (large dots), in each cylinder, as a function of the crankshaft angle. Until a crankshaft angle of 720° in the diagram, the crankshaft rotates a first angular distance of 144° between each ignition of the engine, which distance is the ignition interval at the four stroke mode. After crankshaft angle of 1080° in the diagram, the crankshaft rotates a second angular distance of 216° between each ignition of the engine, which distance is the ignition interval at the six stroke mode. It should be noted that engines with an even number of cylinders, can not alter between four and six stroke modes and present equal distances between the ignitions in both modes.
An ignition interval transition from four stroke to six stroke mode can be initiated at any cylinder, but in this example it is assumed that the transition is initiated at cylinder number 1. This means that after 720 crankshaft degrees, the control device 16 does not send signals so as for the following ignition to take place at 864° in cylinder number 2, as would have been the case in a continued four stroke operation. Instead the control device sends signals so as for the following ignition to take place in cylinder number 4 at 1008°. This means that the crankshaft rotates a third angular distance of 288° between two consecutive ignitions of the engine.
It is not possible to have the ignition after the one at 720° appearing after 216°, due to the requirements mentioned above, restricting the choice of cylinders, in which a following ignition can take place, to those where at least four strokes have been completed after a preceding ignition in the same cylinder. This becomes clear by looking in the diagram in fig. 4. For an ignition to take place at 720°+216°=936°, to meet the TDC condition, the ignition would have to appear in cylinder number 3. However, since the preceding ignition in cylinder 3 took place at 576°, only two strokes (360°) have been completed after the last ignition and a fresh mixture can not be present in the cylinder.
Thus, during the transition, one interval between two ignitions of the engine becomes longer than the normal ignition intervals of four and six stroke operation. This creates a "hole" in the sequence of ignitions. This hole can create a temporary decrease of the engine output torque, and such a decrease could be experienced as unpleasant to the driver and passengers of a vehicle in which the engine is operating. Additionally, the decrease of the engine output torque can be detrimental to the drivability of the vehicle and cause a dangerous situation in the operation of the vehicle.
Preferably, the air and fuel input means 17 are controlled so that the output torque of the engine is essentially continuous during the transition. More specifically, signals are sent from the control device 16 to the air and fuel input means 17 of one of the cylinders of the engine, so as to adjust the combustion load by adjusting the fuel and air input in order to compensate for one of the engine ignition intervals being longer than a normal four or six stroke interval.
A combustion load adjusted to compensate for unequal engine ignition intervals in a transition between two stroke modes is herein also referred to as a transition combustion load.
Preferably, the combustion load is increased at the ignition (at 720° in fig. 4), up to which engine ignition intervals according to the first stroke mode has taken place, and after which the crankshaft rotates the third angular distance (of 288°) between two consecutive ignitions of the engine. In the example in fig. 4, this means that the air and fuel input means 17 of cylinder 1 receives signals so as to increase the injection of fuel in the induction stroke preceding the ignition at 720°. Preferably, the fuel and air input is adapted so that the combustion load is essentially twice as large, compared to the combustion load at the preceding ignition. However, it should be kept in mind that this relative value of the combustion load is theoretical, i.e. no consideration has been made for calibration issues. The reason for the combustion load to be twice as large as the combustion load at the preceding ignition is that the following engine ignition interval is twice as long (288°) as the preceding engine ignition interval (144°). Thereby, the "average output torque" will be continuous since the combustion loads at each ignition is proportional to the length of the engine ignition interval following the ignition. In other words, the output torque of the engine is essentially continuous during the transition.
Alternatively, the combustion load is increased at the ignition occurring at the end of the engine ignition interval of the third angular distance.
As a further alternative, the combustion load at the beginning of the engine ignition interval of the third angular distance, as well as the combustion load at the end of the engine ignition interval of the third angular distance are increased to jointly compensate for the third angular distance being non-equal to the first and second angular distance. Thereby, a compensation factor can be distributed between the two combustion loads in question. This provides for a smoother compensation of the unequalness of the ignition intervals.
In the description above, a transition from four stroke to six stroke mode has been described. However, the invention is equally applicable in transitions from six stroke to four stroke mode, i.e. where the first stroke mode is six stroke mode and the second stroke mode is four stroke mode.
Additionally, the invention is applicable for transitions between higher stroke modes, e.g. from six stroke mode to eight stroke mode or vice versa. An example of eight stroke mode is shown in fig. 5. The ignitions are indicated with large dots, and it can be seen that the interval between each ignition of the engine is 288 degrees of crankshaft angle, which is the length of the eight stroke cycle, 1440°, divided by the number of cylinders: five.
Fig. 6 shows a transition from a six stroke mode to an eight stroke mode. Up until 864 degrees of crankshaft rotation the engine operates in a six stroke mode and after that the eight stroke mode takes over. As opposed to the case of a transition from four to six stroke mode, in the transition from six to eight stroke mode only two engine ignition interval distances occur, one of 216° and the other of 288°.
In each stroke mode the combustion load at each ignition is dependent upon the distance of the ignition interval immediately following or preceding the respective ignition. If a combustion load factor is 1 for combustion loads in six stroke mode, disregarding calibration parameters, a suitable load factor for eight stroke mode would be 1 1/3, since the ignition intervals being 288°/216°=1 1/3 larger in the eight stroke mode.
Referring to fig. 6, if combustion loads at all ignitions up until and including the one at 864° are controlled so that their load factor is 1, and combustion loads of all subsequent ignitions are controlled so that their load factor is 1.33, there will be a momentary drop of the engine torque output in the transition. The reason is that the combustion load at the ignition at 864° of crankshaft angle rotation is too small to compensate for the larger ignition interval immediately following the ignition. This could affect the comfort of the driver and passengers of a vehicle in which the engine is operating, and could also affect the driveability of the vehicle.
Similarly, still referring to fig. 6, if combustion loads at all ignitions up until and including the one at 648° are controlled so that their load factor is 1, and combustion loads of all subsequent ignitions are controlled so that their load factor is 1.33, there will be a momentary increase of the engine torque output in the transition. The reason is that the combustion load at the ignition at 864° of crankshaft angle rotation is too large, since the interval between 648° and 864° of crankshaft angle rotation is smaller than the consecutive ones. As in the case with a drop in the output torque the comfort of the driver and passengers and the driveability of a vehicle in which the engine is operating could be affected.
According to a special embodiment of the invention, the combustion load related to the ignition at 864° is determined at least partly on the basis of the ignition intervals before and after the ignition, i.e. 216° and 288°. In general, one of the transition combustion loads, related to an ignition at which the crankshaft rotates essentially a first angular distance between the ignition and an immediate preceding ignition, and the crankshaft rotates essentially the second angular distance between the ignition and an immediate following ignition, is determined at least partly on the basis of the first and the second angular distance.
Preferably, the magnitude of the transition combustion load is between the magnitudes of the combustion loads of the preceding and following ignitions. In the example in fig. 6 this means that the combustion load at 864° gets a load factor between 1 and 1 1/3. Suitably, the load factor at 864° is determined as (1+(1 1/3))/2 = 1 1/6, i.e. the mean value of load factors for preceding and following ignitions.
The method described with reference to fig. 6 can be applied in any transition from any stroke mode to any other, in engines with any number of cylinders. In particular, in a case where a transition is made from eight to six stroke mode, and also in cases of transitions from four stroke to eight stroke mode, and vice versa, the method described with reference to fig. 6 can be applied.

Claims

A method in an internal combustion engine including a crankshaft, whereby controllable valves and controllable air and fuel input means are provided at each cylinder, the method comprising the steps of

controlling the valves and the air and fuel input means so that the engine is operating in a first stroke mode, in which the crankshaft rotates essentially a first angular distance between two consecutive ignitions of the engine, and

controlling the valves and the air and fuel input means so that the engine is operating in a second stroke mode, in which the crankshaft rotates essentially a second angular distance between two consecutive ignitions of the engine, the second angular distance being non-equal to the first angular distance,

characterized by

controlling the valves and the air and fuel input means so that a transition is made from the first stroke mode to the second stroke mode, and

controlling at least one transition combustion load at or near the transition, the transition combustion load being determined at least partly on the basis of the first and/or the second angular distance.
A method according to claim 1, whereby, in the transition, the crankshaft rotates essentially a third angular distance between two consecutive ignitions of the engine, the third angular distance being non-equal to the first and the second angular distances, the transition combustion load being determined at least partly on the basis of the third angular distance in relation to the first angular distance and/or the second angular distance.
A method according to claim 2, wherein the transition combustion load is related to an ignition occurring at the beginning or the end of the third angular distance rotation of the crankshaft.
A method according to claim 3, wherein the transition combustion load is related to an ignition occurring at the beginning of the third angular distance rotation, and is determined at least partly on the basis of the third angular distance in relation to the first angular distance.
A method according to claim 3, wherein the transition combustion load is related to an ignition occurring at the end of the third angular distance rotation, and is determined at least partly on the basis of the third angular distance in relation to the second angular distance.
A method according to claim 2, wherein the transition combustion load is related to an ignition occurring at the beginning of the third angular distance rotation of the crankshaft, and a further transition combustion load is related to an ignition occurring at the end of the third angular distance rotation of the crankshaft.
A method according to claim 1, whereby one of the transition combustion loads is related to an ignition at which the crankshaft rotates essentially the first angular distance between the ignition and an immediate preceding ignition, and the crankshaft rotates essentially the second angular distance between the ignition and an immediate following ignition, the transition combustion load being determined at least partly on the basis of the first and the second angular distance.
A method according to claim 7, whereby the magnitude of the transition combustion load is between the magnitudes of the combustion loads of the preceding and following ignitions.