EP2718556B1 - Method for controlling rail pressure - Google Patents

Description

The invention relates to a method for regulating the rail pressure of an internal combustion engine in V-arrangement with unequal ignition sequence according to the preamble of claim 1.

Internal combustion engines in V arrangement have on the A and the B side a rail for intermediate storage of the fuel. Connected to the rail are the injectors, via which the fuel is injected into the combustion chambers. In a first design of the common rail system, a single high-pressure pump delivers the fuel in parallel pressure in both rails. There is therefore the same rail pressure in both rails. A second design of the common rail system differs in that a first high pressure pump in a first rail and a second high-pressure pump in a second rail promote. Both types are for example from the DE 43 35 171 C1 known.

Since the quality of the combustion depends crucially on the pressure level in the rail, this is regulated. Typically, a rail pressure control loop comprises a pressure regulator, the suction throttle with high pressure pump and the rail as a controlled system and a software filter in the feedback branch. In this rail pressure control loop, the pressure level in the rail corresponds to the controlled variable. The measured raw values of the rail pressure are converted via the filter into an actual rail pressure and compared with a desired rail pressure. The resulting deviation is then converted via the pressure regulator into a control signal for the suction throttle. The control signal corresponds to a volume flow with the unit liters / minute, which is electrically designed as a PWM signal (pulse width modulated). A corresponding rail pressure control loop is from the DE 10 2006 049 266 B3 known.

From the DE 10 2007 034 317 A1 is an internal combustion engine in V-arrangement with unequal firing order and a separate A-side and an independent B-side common rail system known. An unequal firing order is given when, for example, the cylinder A1, ie the first cylinder on the A side, is ignited and, following the cylinder A2, that is the second cylinder on the A side, ignited. The unequal ignition sequence in turn causes pressure fluctuations in the rail. To solve this problem proposes the DE 10 2007 034 317 A1 in a first solution before a compensation line between the two rails. In a second solution, the A-side rail pressure in an A-side rail pressure control loop is controlled with a PI controller and the B-side rail pressure in a B-side rail pressure control loop with a P-control. Due to the missing B-side I-component of the controller, this solution is critical in terms of a permanent control deviation.

The invention is therefore based on the object to design an improved rail pressure control in an internal combustion engine in V-arrangement with an unequal firing order.

The invention solves this problem by a method for rail pressure control with the features of claim 1. The embodiments are shown in the subclaims.

According to the invention, the actual rail pressure is calculated via an average filter from the measured rail pressure by averaging the rail pressure below a limit speed over a constant time and averaged over the limit speed of the rail pressure over a cycle of the internal combustion engine. By working cycle are meant two revolutions of the crankshaft. This solution has proven particularly useful in an internal combustion engine generator application in which the engine speed passes through different engine speed ranges during engine operation. In the stationary speed range, for example at a constant engine speed of 1500 rpm for generating a 50 Hz power frequency, the working cycle periodic rail pressure oscillations are filtered out by averaging the rail pressure over a working cycle of the internal combustion engine. In a speed range below the stationary speed range, for example, from zero revolutions to a limit speed of 1000 revolutions per minute, the rail pressure, however, is averaged over a constant time. This measure causes the signal of the actual rail pressure is not delayed too much below the limit speed, which in turn allows only a satisfactory control of the rail pressure. An advantage is therefore a stabilization of the rail pressure control loop below the limit speed.

In an internal combustion engine generator application, this ensures that an averaging of the rail pressure via a working cycle reliably takes place in the stationarily operable operating range, since the rail pressure oscillations are working cycle-periodic. In the speed range below the limit speed, however, an exact averaging over a working cycle and thus an exact filtering out the working cycle periodic rail pressure vibrations is not required because the area below the limit speed is only dynamically traversed and therefore rail pressure oscillations can develop here not sustainable.

In one embodiment, the mean value filter is combined with a low-pass filter, as a result of which high-frequency rail pressure oscillations, which are not periodic in the working cycle, are damped.

The method can be used both in an internal combustion engine in V-arrangement with unequal firing order and with a separate A-side and an independent B-side common rail system as well as in an internal combustion engine in V arrangement with unequal firing order, in which a single high-pressure pump Promotes the fuel at the same time in the A-side and the B-side rail.

Figur 1

ein Systemschaubild,

Figur 2

ein Blockschaltbild des Raildruck-Regelkreises,

Figur 3

eine Kennlinie,

Figur 4

ein Zeitdiagramm und

Figur 5

einen Programm-Ablaufplan.

In the figures, a preferred embodiment is shown. Show it:

FIG. 1

a system diagram,

FIG. 2

a block diagram of the rail pressure control loop,

FIG. 3

a characteristic

FIG. 4

a time chart and

FIG. 5

a program schedule.

The FIG. 1 shows a system diagram of an electronically controlled internal combustion engine 1 with a common rail system on the A side and a common rail system on the B side. The common rail system on the A side comprises as mechanical components a low-pressure pump 3A for conveying fuel from a tank 2, a suction throttle 4A for influencing the volume flow, a high-pressure pump 5A, a rail 6A and injectors 7A for Injecting fuel into the combustion chambers of the internal combustion engine 1. The common rail system on the B side comprises the same mechanical components, which are indicated by the suffix B at the reference numerals.

The internal combustion engine 1 is controlled via an electronic engine control unit 10 (ECU). In the FIG. 1 For example, an A-side rail pressure pCR (A), a B-side rail pressure pCR (B), and a size ON are shown as inputs to the electronic engine control unit 10. The A-side rail pressure pCR (A) is detected via an A-side rail pressure sensor 9A. The B-side rail pressure pCR (B) is detected via a B-side rail pressure sensor 9B. The size ON is representative of the other input signals, for example for an engine speed or for a performance request of the operator. The illustrated outputs of the electronic engine control unit 10 are a PWM signal SD (A) for driving the A-side intake throttle 4A, a power-determining signal ve (A) for driving the A-side injectors 7A, for example injection start / end injection, a PWM Signal SD (B) for driving the B-side suction throttle 4B, a power-determining signal ve (B) for driving the B-side injectors 7B and a size OFF. The latter is representative of the other control signals for controlling the internal combustion engine 1, for example, a control signal for controlling an EGR valve. The illustrated common rail system can also be designed as a common rail system with individual memories. In this case, then in the injector 7A a single memory 8A and in the injector 7B a single memory 8B are integrated as additional buffer volumes for the fuel. The individual storage pressures pE (A) and pE (B) are then further inputs of the electronic engine control unit 10. Characteristic feature of the illustrated embodiment is the independent control of the A-side rail pressure pCR (A) and the independent control of the B-side rail pressure pCR (B).

The FIG. 2 shows a block diagram of the A-side rail pressure control loop, which is indicated in the figure by the addition A at the reference numerals. Both control circuits are identical. In the following, the A-side rail pressure control loop 11A will be described, the description of which applies mutatis mutandis to the B-side rail pressure control loop. The reference variable is identical for both rail pressure control circuits, here: a common target rail pressure pCR (SL). The target rail pressure is calculated as a function of a desired torque or as a function of the desired injection quantity and the engine speed.

The input variables of the rail pressure control loop 11A are the target rail pressure pCR (SL), a basic frequency fPWM for the PWM signal, a quantity E1, the engine speed nMOT, a time constant T1 and a time constant T2. The input quantity E1 comprises the battery voltage and the resistance of the suction throttle including the supply line, which are included in the calculation of the drive signal SD (A) for the suction throttle 4A. The output of the A-side rail pressure control loop are the raw values of the rail pressure pCR (A). The raw values of the rail pressure pCR (A) are measured by the A-side rail pressure sensor 9A. Its output signal pMESS is then filtered via a hardware filter 16A with PT1 behavior and a corner frequency of 20 Hz. The output values pHW are digitized by an A / D converter 17A. The output values pAD of the A / D converter 17A are then further processed via two information paths. A first information path includes a mean value filter 18A and an optional low pass filter 19A. The first information path corresponds to a slow filtering, via which the actual rail pressure pIST (A) is determined. The mean value filter 18A has, as further input variables, the engine speed nMOT and the limit speed nLi. Via the mean value filter 18A it is determined whether the averaging of the rail pressure takes place either via a working cycle, ie two revolutions of the crankshaft, or over a constant time. The switching between the two methods of averaging takes place at the limit speed nLi. The output pMW of the average filter 18A is then further processed by the low-pass filter 19A, as shown. This has a time constant T1 as input. In practice, a value of T1 = 16 ms is used for the time constant, which corresponds to a frequency of 10 Hz. About the low-pass filter 19A high-frequency rail pressure vibrations, which are not Arbeitsspielperiodisch attenuated. A second information path includes a fast filter 20A with PT1 behavior. The fast filter 20A in this case has a smaller time constant and thus a lower phase delay than the average value filter 18A and the optional low-pass filter 19A. The output value pDYN (A) of the fast filter 20A is used, inter alia, to perform a rapid energization of the suction throttle, whereby a higher dynamics is achieved in a load shedding.

At a point A, the actual rail pressure pIST (A) is compared with the target rail pressure pCR (SL). This results in the control deviation ep (A), from which a pressure regulator 12A with at least PID behavior calculates a setpoint volume flow VSL as a manipulated variable. The nominal volume flow VSL has the physical unit liters / minute. Thereafter, the desired volume flow is limited (not shown) and the desired volume flow VSL via a pump characteristic 13A assigned a target electric current iSL. The target current iSL is converted in a calculation 14A into a PWM signal SD (A). The PWM signal SD (A) is the duty cycle and the frequency fPWM corresponds to the fundamental frequency of the PWM signal SD (A). During the conversion, inter alia, the fluctuations of the operating voltage and the ohmic resistance of the suction throttle including the electrical leads are taken into account. With the PWM signal SD (A) then the solenoid of the A-side suction throttle is applied. As a result, the path of the magnetic core is changed, whereby the flow rate of the high-pressure pump is influenced freely. The high-pressure pump 5A, the suction throttle 4A and the rail 6A correspond to an A-side controlled system 15A. Thus, the A-side control loop 11A is closed.

The FIG. 3 shows a characteristic curve 21. On the characteristic curve 21, an averaging time dT is calculated as a function of the engine speed nMOT. The averaging time dT thus corresponds to the time over which the rail pressure values from the mean value filter ( Fig. 2 : 18A) are averaged. The characteristic curve 21 is composed of an abscissa-parallel straight line 22 and a hyperbola 23. For engine speed values smaller than a limit speed nLi = 1000 rpm, a constant averaging time dT = 120 ms is determined via the straight line 22. This area is in the FIG. 3 hatched shown. The averaging time dT = 120 ms is calculated from the duration of a working cycle at a speed of 1000 rpm. A working cycle corresponds to two revolutions of the crankshaft of the internal combustion engine, ie 720 ° crankshaft angle. Below the limit speed nLi, the rail pressure is filtered with a constant averaging time dT = 120 ms. For larger engine speed values nMOT than the limit speed nLi = 1000 1 / min, the averaging time dT corresponds to a working cycle, which gives the hyperbola 23. For example, at an engine speed nMOT = 1500 1 / min, an averaging time dT = 80 ms or at an engine speed nMOT = 2000 1 / min an averaging time of dT = 60 ms is calculated.

The FIG. 4 consists of the subfigures 4A to 4C, which show different state variables. Over time t are shown: the engine speed nMOT in FIG. 4A , the averaging time dT in FIG. 4B and the average rail pressure pMW in FIG. 4C ,

In the FIG. 4A the startup procedure and a load application in an internal combustion engine generator arrangement are shown. The target speed nSL is shown as a dot-dash line and the limit speed nLi as a dashed line in the FIG. 4A located. The setpoint speed remains constant at nSL = 1500 1 / min, which corresponds to a frequency of 50 Hz. The engine speed nMOT reaches the limit speed of nLi = 1000 1 / min at time t1. At time t2, the target speed of nSL = 1500 1 / min is reached. After a speed overshoot, the engine speed nMOT is steady at the target speed nSL at time t4. At time t6, a load application takes place, which leads to a collapse of the engine speed nMOT. In the period t7 to t8, the engine speed falls below the limit speed nLi. Due to the target actual deviation of the engine speed now more fuel is injected, whereby the engine speed nMOT increases again. At the time t9, the engine speed nMOT again reaches the speed level of the setpoint speed nSL and has settled at the setpoint speed nSL at the time t10.

The FIG. 4B shows the averaging time dT, over which the rail pressure values, for example the A-side rail pressure pCR (A), are averaged. Until the time t1, the engine speed nMOT is smaller than the limit speed nLi. About the characteristic of the FIG. 3 Therefore, a constant averaging time dT = 120 ms is calculated. In the speed range below the limit speed nLi an exact averaging over a working cycle is not required because this area is only passed through dynamically and therefore do not develop rail pressure vibrations here. The averaging over a constant time has a stabilizing effect on the rail pressure control since the signal of the actual rail pressure is not delayed too much. After the time t1, the engine speed nMOT is greater than the limit speed nLi. Now the averaging time dT is calculated as a function of the engine speed nMOT via the hyperbola of the FIG. 3 , As a result, the averaging time dT decreases as the engine speed nMOT increases. Since the rail pressure is now averaged over a cycle of the internal combustion engine, the working cycle periodic fluctuations of the rail pressure are filtered out.

At the time t4, the engine speed nMOT has settled at the target speed nSL = 1500 rpm. This means that the averaging time has stabilized at the value dT = 80 ms. If a load connection occurs at time t6, the averaging time dT increases due to the decreasing engine speed. In the period t7 / t8, the engine speed falls below the limit speed nLi = 1000 1 / min. Now it is about the characteristic of the FIG. 3 , here: Just 22, calculated a constant averaging time of dT = 120 ms. From the time t8, the engine speed nMOT again exceeds the limit speed nLi, so that the averaging time now again as a function of the engine speed ( Fig. 3 : Hyperbola 23) is calculated.

The diagram of FIG. 4C shows the averaged rail pressure pMW, which initially increases and at time t3 reaches the constant target rail pressure pCR (SL) = 800 bar. After an overshoot, the averaged rail pressure pMW settles at the desired rail pressure pCR (SL) at time t5. As shown, the speed drop due to the load application has only a minor effect on the average rail pressure pMW.

In the FIG. 5 the procedure is shown in a program flowchart as a subroutine. At S1 it is checked whether the engine speed nMOT is greater than or equal to the limit speed nLi. In practice nLi = 1000 1 / min is chosen. If the engine speed nMOT is above the limit speed nLi, query result S1: yes, then at S2 the number of values N, via which the rail pressure is averaged, is calculated according to the engine speed nMOT and the sampling time tS. For nMOT = 1500 1 / min and a sampling time tS = 1 ms this gives a number of N = 80 values. If the engine speed nMOT is smaller than the limit speed nLi, query result S1: no, then the number N is not calculated in S3 as a function of the engine speed nMOT but based on the constant preset limit speed nLi. For a limit speed nLi = 1000 1 / min these are N = 120 values. Then the program schedule is finished.

reference numeral

1

Internal combustion engine

2

tank

3A, 3B

Low pressure pump

4A, 4B

interphase

5A, 5B

high pressure pump

6A, 6B

Rail

7A, 7B

injector

8A, 8B

Single memory (optional)

9A, 9B

Rail pressure sensor

10

electronic engine control unit (ECU)

11A, 11B

Rail pressure control circuit

12A, 12B

pressure regulator

13A, 13B

Pump curve

14A, 14B

Calculation PWM signal

15A, 15B

controlled system

16A, 16B

hardware filters

17A, 17B

A / D converter

18A, 18B

Average filter

19A, 19B

Low pass filter (PT1 filter, optional)

20A, 20B

fast filter (PT1 filter)

21

curve

22

Just

23

hyperbole

Claims (4)

1. Method for rail pressure regulation in an internal combustion engine (1) in a V arrangement with nonuniform ignition sequence, in which method an actual rail pressure is calculated from the measured rail pressure, a regulation deviation is determined on the basis of the actual rail pressure and a setpoint rail pressure, and in which method an actuation variable for the activation of a pressure actuator, in particular of a suction throttle, for the regulation of the rail pressure is calculated,
   characterized
   in that the actual rail pressure (pIST(A), pIST(B)) is calculated by means of an average-value filter (18A, 18B) from the measured rail pressure (pCR(A), pCR(B)) in that, below a threshold rotational speed (nLi), the rail pressure (pCR(A), pCR(B)) is averaged over a constant time and, above the threshold rotational speed (nLi), the rail pressure (pCR(A), pCR(B)) is averaged over one working cycle of the internal combustion engine (1).
2. Method according to Claim 1,
   characterized
   in that, in addition, the actual rail pressure (pIST(A), pIST(B)) is calculated by means of a low-pass filter (19A, 19B).
3. Method according to Claim 1,
   characterized
   in that the rail pressure (pCR(A)) of the common rail system on the A side and the rail pressure (pCR(B)) of the common rail system on the B side are regulated independently of one another by means of an A-side rail pressure regulation loop (11A) and by means of a B-side rail pressure regulation loop (11B) respectively, and a common setpoint rail pressure (pSL) is set as a reference variable for both rail pressure regulation loops (11A, 11B).
4. Method according to one of the preceding claims,
   characterized
   in that the common setpoint rail pressure (pCR(SL)) is calculated as a function of a setpoint torque or as a function of the setpoint injection quantity and the engine speed (nMOT).

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