Monitoring of diesel engine combustions based on the acoustic source characterisation of the exhaust system

ARTICLE IN PRESS Mechanical Systems and Signal Processing Mechanical Systems and Signal Processing 22 (2008) 1465–1480 www.elsevier.com/locate/jnlabr/ymssp

Monitoring of diesel engine combustions based on the acoustic source characterisation of the exhaust system J. Jianga, F. Gub,, R. Gennisha, D.J. Moorea, G. Harrisa, A.D. Ballb a

School of Mechanical, Aerospace and Civil Engineering, University of Manchester, Manchester M13 9PL, UK b School of Computing and Engineering, University of Huddersfield, Queensgate, Huddersfield HD1 3DH, UK Received 17 July 2007; received in revised form 7 October 2007; accepted 16 December 2007 Available online 31 December 2007

Abstract Acoustic methods are among the most useful techniques for monitoring the condition of machines. However, the influence of background noise is a major issue in implementing this method. This paper introduces an effective monitoring approach to diesel engine combustion based on acoustic one-port source theory and exhaust acoustic measurements. It has been found that the strength, in terms of pressure, of the engine acoustic source is able to provide a more accurate representation of the engine combustion because it is obtained by minimising the reflection effects in the exhaust system. A multi-load acoustic method was then developed to determine the pressure signal when a four-cylinder diesel engine was tested with faults in the fuel injector and exhaust valve. From the experimental results, it is shown that a two-load acoustic method is sufficient to permit the detection and diagnosis of abnormalities in the pressure signal, caused by the faults. This then provides a novel and yet reliable method to achieve condition monitoring of diesel engines even if they operate in high noise environments such as standby power stations and vessel chambers. Crown Copyright r 2007 Published by Elsevier Ltd. All rights reserved. Keywords: Condition monitoring; Combustion diagnosis; One-port source; Exhaust acoustics; Diesel engine

1. Introduction The exhaust stream is directly related to the combustion process and contains rich information about engine combustion conditions. Recently, with growing interest in condition monitoring, more and more advanced techniques are being introduced to detect and diagnose faults in internal combustion (IC) engines based on the information from the exhaust stream, such as the exhaust gas temperature techniques [1], turbocharger speed measurements [2,3] and exhaust pressure fluctuations [4–6]. It seems that the pressure fluctuation methods have shown very promising results for the combustion diagnosis in regards to misfire detection. However, there inevitably exists the issue of acoustic reflection and standing wave effects inside the exhaust pipe. The results obtained directly from the output of the pressure sensor may be not so reliable [7] because the measured pressures vary with the sensor positions along the pipe. In essence, an abnormal combustion may be Corresponding author. Tel.: +44 161 2754458; fax: +44 161 2755441.

E-mail address: [email protected] (F. Gu). 0888-3270/$ - see front matter Crown Copyright r 2007 Published by Elsevier Ltd. All rights reserved. doi:10.1016/j.ymssp.2007.12.003

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detected by positioning the sensor in one location but may not be detected at another location. To improve the reliability, the standing wave effects have to be considered and a measurement of the exhaust source characteristics has to be established that is independent of the reflections from the downstream components. The establishment of such a measurement has been studied extensively in IC engine exhaust system development. For the effective design of a silencer system and engine intake and exhaust processes, the understanding of sound generation mechanisms and the quantification of the source are often based on acoustic one-port models. In particular, the exhaust source (i.e. engine combustion and exhaust process) is assumed to be a linear time-invariant source in the frequency domain and can be fully described by two characteristic parameters: source strength and source impedance without any influences from the exhaust systems connecting to the engine [8–10]. Based on the acoustic one-port model, various experimental methods have been developed to determine the source characteristics, including the standing wave method, transfer function method and multi-load method. They were developed initially for silencer designs [11–13]. Due to the impressive performance of acoustic parameters in describing characteristics of acoustic sources, a considerable amount of work has been carried out on this subject but with various fluid machines such as pumps, fans and engines. Ross and Crocker [14] measured the acoustic impedance of an eight-cylinder engine. The recorded signal was processed to obtain standing waves at various frequencies. Both the speed and the load of the engine were considered to be the only factors influencing the engine condition. They observed that at the higher acoustic frequencies, the engine’s internal impedance tends to approach the characteristic impedance. Doige and Thawani [15] used the standing wave technique to measure the source impedance of a nonoperating compressor. The convective effects were considered as a correction of the impedance. A method of the correction was given by Munjal [16] and hence a good agreement was obtained. Prasad and Crocker [17] described the use of a transfer function method (with a random excitation source) for measurement of the internal source impedance of an eight-cylinder engine under running conditions. The results obtained agree well with those obtained from the standing wave method by earlier investigators. Among these methods, the multi-load method is of the most interesting. In particular, Boden et al. [8,9,18] published a series of work that discussed the multi-load methods used for determining the impedance of the source for a one-port system, including the two-load method, four-load method and associated direct least-square method. By comparing the results derived from different methods, Boden [18] concluded that the two-load method should be considered first when there was no external source used in the experiment, since it is less sensitive to errors in the input data than the other methods. In addition to obtaining good results, the two-load method can also be implemented easily on-line as only two acoustic load conditions are required to connect to the engine exhaust system. Therefore, this paper focuses on the use of the two-load method for the development of combustion monitoring techniques. The acoustic source parameters from a four-cylinder diesel engine are studied in line with the engine combustion process to find the important features for combustion diagnosis. The monitoring performance using these features is evaluated by monitoring different abnormal combustions due to common faults such as abnormal fuel injection and incorrect valve timing. 2. Acoustic model of engine combustion and exhaust system For the purpose of noise control and flow optimisation, the acoustic one-port source method has been widely used to describe characteristics of acoustic sources in fluid machines [8] including IC engine exhaust systems [10,17]. To develop engine combustion monitoring methods based on this source model, the theoretical basis behind it is overviewed in this section so that an experimental determination of the source characteristics can be developed with a two-sensor and two-load approach and fault features can be extracted from the source data. 2.1. Acoustic model To develop an easy procedure for the design of an exhaust system, an acoustic one-port source was introduced in Refs. [17,20] to avoid the complexity of fluid dynamic processes through the exhaust valves in

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Source Zs , Ps

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v Sound wave P, v

Silencer Zl Zs

P

Z

Ps

Fig. 1. Acoustic source and load model for an engine exhaust system. (a) Schematic of engine combustion and silencer, (b) analogous electro acoustic circuit.

acoustic calculations. As shown in Fig. 1(a), it treats the engine acoustic sources including the combustion and its exhaust process with two parameters: source strength Ps and source impedance Zs. All the components downstream of exhaust valves are regarded as an acoustic load of the impedance Zl applying to the engine source. Based on these simplifications, the complicated combustion and exhaust acoustics can be represented by an equivalent electric circuit as shown in Fig. 1(b), and the acoustic pressure at the source and load interface can be simply expressed using the source parameters P¼

Ps Z l . Zs þ Zl

(1)

Based on this relationship, acoustic performances such as radiated acoustic power and insert loss of a silencer can be predicted easily, simplifying the design and assessment of exhaust systems. Although this model is a linear time-invariant source representation, the predicted results show good qualitative agreement with tests [21]. Interestingly, Eq. (1) also shows that the source parameters can be expressed by the acoustic load parameter downstream. This means that an appropriate acoustic measurement from the exhaust system can be used to predict the combustion source characteristics which are difficult to obtain by a direct measurement. These predicted source parameters can then be used to achieve condition monitoring of the engine combustion. To achieve the exhaust acoustics based on the combustion monitoring scheme, the physical means behind the acoustic parameters need to be further explored. The studies based on the basic fluid dynamic equations for inviscid one-dimensional flow made by Dokumaci [22] recently have shown that the acoustic source parameters connect closely to the engine combustion process: Ps 

g  1 ðmhÞ0 , g S¯v

(2)

p¯ (3) Zs  , v¯ where g is the ratio of specific heat coefficients, m is the rate of exhaust gas mass injecting through the valve port, S is the cross-sectional area of valve, h is the specific enthalpy of the gas in the cylinder, p¯ and v¯ denote the mean pressure and the mean flow velocity, respectively, just downstream of the source discontinuity plane (exhaust valve). From Eqs. (2) and (3), it is known that the source strength is proportional to the fluctuations of the stagnation enthalpy and is inversely proportional to the mean flow velocity. On the other hand, the source impedance depends only on the mean pressure and velocity. This is consistent with the experimental observations that acoustic waves in the exhaust stream are closely correlated with engine combustion and exhaust system. A high intensity of combustion usually results in a high amplitude acoustic wave in the exhaust pipe. A different profile of the exhaust valve when opening and closing will lead to changes of the gas flow and hence the acoustic waveforms. In addition, an inadequate profile will cause poor performance in the engine exhaust stroke, which in turn will influence the combustion efficiency. These relationships provide the

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basis for the feasibility of acoustic exhaust-based monitoring. However, the acoustics are influenced significantly by acoustic reflections from the downstream components such as silencers, pipeline junctions, etc. Therefore, suppressing the reflection and hence determining the true acoustic source is the key task in developing the monitoring scheme. On the other hand, the assumptions, linear and time-invariant, seem not the realities of an engine. In most cases, the engine emits acoustic waveforms with high amplitudes. Further superposed by the reflection, the overall acoustic waves may reach a nonlinear level easily. This will cause the frequency shift of the waveform during the propagation. In addition, the acoustic signal emitted by the engine is not strictly time-invariant because the operating process of the exhaust valve is highly time-dependent. Several studies [21,23] have been conducted recently to explore this subject and some improvements on the results have been obtained. However, taking into account these two aspects would make the calculation procedure very complex. For a condition monitoring purpose, the methods developed must be easy and simple to implement on-line or in the field. Therefore, investigation into the possibility of using a linear and time-invariant model to define the engine system is preferred in achieving an efficient condition monitoring system. In addition, the detection of slight faults is another task that needs to be carried out. In most previous studies, the frequency domain analysis is sufficient. However, this method cannot perform very well with small fault signals with random noise. In contrast to this, a time domain waveform analysis is supposed to achieve this. Therefore, acoustic waveforms with different faults in various conditions will be carefully studied in this paper. 2.2. Determination of the source The previous work and the measurement methods available for estimation of the characteristics of a linear one-port source are encompassed in the review article by Boden [18]. They consider the existing measurement methods in two groups. The methods in the first group require the use of an external sound source, whereas those in the second group require the use of external loads. The present paper will contribute to the second group of methods when complex pressure measurements can be made at the source plane. In this case, if the equivalent source is time-invariant, the one-port source characteristics can be determined by using only two external loads. This is known as the two-load method. As shown in Fig. 2, the diesel engine combustion and exhaust system can be modelled as an acoustic system that has a long straight pipe connecting the source (engine and its manifold) and the downstream silencer. In Fig. 2, position 0 refers to the reference plane; positions 1 and 2 are the two locations at which two pressure transducers are mounted. PI and PR represent the pressure waves propagating forward (away from the engine) and backward (reflected from the muffler), respectively. For plane wave propagation, the forward and backward waves PI and PR are represented as ( PI ¼ Pi ejðotkxÞ , (4) PR ¼ Pr ejðotþkxÞ pffiffiffiffiffiffiffi where o is angular frequency, k ¼ o/c is wave number, t is time, x is the distance and j ¼ 1.

1

2 PI

PR

d

Silencer

Δl 0

l1

l2

x

Fig. 2. A schematic of the diesel engine and exhaust system.

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Then, the volumetric forward and backward velocities, vI and vR, can be expressed, respectively, as 8 PI Pi jðotkxÞ > > > < vI ¼ r c0 ¼ r c0 e 0 0 P Pr jðotþkxÞ . R > > v ¼  ¼  e > R : r0 c0 r0 c0

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(5)

As the pressure measured at the position 1 is P1 ¼ PI1 þ PR1 ¼ Pi ejðotkl 1 Þ þ Pr ejðotþkl 1 Þ ,

(6)

and at the position 2 is P2 ¼ PI2 þ PR2 ¼ Pi ejðotkl 2 Þ þ Pr ejðotþkl 2 Þ .

(7)

By solving the set of Eqs. (6) and (7), the forward and backward waves Pi and Pr can be expressed by pressures P1 and P2, and sensor locations l1 and l2: 8 P1 ejkl 2  P2 ejkl 1 jot > > e P ¼  > i < 2j sin kðl 1  l 2 Þ . (8) > P1 ejkl 2  P2 ejkl 1 jot > > e P ¼ r : 2j sin kðl 1  l 2 Þ Hence, the pressure P0 at the reference plane x ¼ 0 can be expressed as P0 ¼ Pi ejot þ Pr ejot ¼

P1 sin kl 2 þ P2 sin kl 1 . sin kðl 1  l 2 Þ

(9)

Volume velocity for the that position is v0 ¼

Pi jot Pr jot P1 cos kl 2 þ P2 cos kl 1 . e  e ¼ r 0 c0 r 0 c0 jr0 c0 sin kðl 1  l 2 Þ

(10)

Then, the reference impedance can be obtained: Z0 ¼

P0 r c0 P1 sin kl 2  P2 sin kl 1 ¼j 0  . sv0 s P1 cos kl 2  P2 cos kl 1

(11)

It should be noted that P0 and v0 in Eqs. (10) and (11) are just the reference parameters and cannot be used directly to describe the sound source directly. From the equivalent circuit in Fig. 1, the pressure and acoustic impedance of the sound source, Ps and Zs can be obtained by P0 and v0 as P0 þ sv0 Z s ¼ Ps .

(12)

As there are two unknowns Ps and Zs in Eq. (12), two equations are required to determine them, provided that the source remains constant. This means two different acoustic loads have to be applied to the source to constitute two reference conditions: I and II as ( ð1Þ PI þ svI Z s ¼ Ps : (13) PII þ svII Zs ¼ Ps That is also the reason why the method is referred to as the two-load method. From Eq. (13), Ps and Zs can be derived as 8 vII PI  vI PII > > < Ps ¼ v  v II I (14) : PI  PII > > : Zs ¼ ðvII  vI Þs Eq. (14) shows that to obtain the source impedance and pressure, two different acoustic loads have to be used. This can be achieved in many ways such as using two different silencers or a variable length of exhaust

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ducts. However, according to error analysis results [19], the determination of Zs and Ps can produce a small error, less than 1%, when one of the load impedances is much smaller or much larger than the other one. For easy implementation, a simple throttling valve may be used to adjust the exhaust flow slightly, which is detailed in the following section. In addition, the calculations of the source impedance and pressure are conducted in the frequency domain for computational efficiency. To obtain them, the measured pressure signals in the time domain have to be converted into the frequency domain first by a forward Fourier transform. However, it is difficult to keep the condition of engine constant when the acoustic load is changed. This will cause the abridgement of waveform if the time interval is strict. The abridgement will then lead to an error in the FFT. To avoid this error, an algorithm which adjusts the integrated waveform of one revolution into a standard time period is used. 3. Experiment techniques 3.1. Test system All the experiments and data used in this work have been collected from a FSD 425 four-cylinder 2.5 l direct injection diesel engine. The major specifications of the engine are summarised in Table 1. The engine is connected to a hydraulic dynamometer so that different loads can be applied to it for the study of acoustic characteristics. To implement the two-load acoustic impedance measurement, a special exhaust system is fitted to the engine. As shown in Figs. 3 and 4, the exhaust gas can flow out through two different exhaust systems by a Y junction fitted with two throttle valves. This configuration allows the application of at least three different acoustic loads to the engine source, which is achieved by adjusting the valves into one of three configurations: (1) both valves are open; (2) valve A is open with valve B closed; and (3) valve B is open with valve A closed. These three configurations can be arranged into three combinations for impedance calculation by coupling any of the two configurations. It was found that the results obtained by using the data from any of the three Table 1 Test engine specifications 2496 mm3 1-2-4-3 250 bar 93.67 mm 90.54 mm 19:1 511BBDC 131ATDC 131BTDC 391ABDC

Capacity Injection sequence Fuel injection pressure Bore Stroke Compression ratio Exhaust valve open Exhaust valve close Inlet valve open Inlet valve close

P2

P1

Fig. 3. Position of pressure transducers on the engine exhaust pipe.

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A

B

Fig. 4. Acoustic load control valves.

combinations are close to each other. Therefore, only the combination between configurations 2 and 3 is used for the impedance calculation. Two pressure transducers and temperature sensors are mounted at two different positions along the straight portion of the exhaust pipe shown in Fig. 3. The measured waves from these positions will have a minimal influence by the differences of pipe shapes. The distance between the two pressure sensors is 320 mm. This distance ensures that the calculation is accurate in a frequency range between 20 and 800 Hz, which covers the frequencies of interest. In addition, a pressure transducer is mounted inside cylinder number 1 to measure the difference in combustion when different faults are introduced. The results from this pressure analysis are then used to benchmark that of the two-load approach. A top dead centre (TDC) trigger signal is used to set the start time of data collection so that each data segment is measured at a specified crank position. This is to ensure accurate time domain averaging and rearrangement of data segments. 3.2. Test procedure Two sets of measurements were conducted. The first set is for the exploration of the characteristics of the engine source at different operating conditions when the engine is in a healthy condition. This involves testing at three different engine loads: 20, 40 and 60 N m at three different engine speeds: 1500, 1800 and 2100 rpm. The second set of measurements is for the detection capability study. A small fault was introduced to the fuel supply system of the engine by changing the fuel injector opening pressure from 250 to 150, 190 and 280 bar, respectively. This is a common fault occurring in diesel engines and causes power loss and high emissions. 4. Results and discussions 4.1. Measured characteristics of the source From the test data, the characteristics of the engine exhaust source are examined in both the time and the frequency domains. Figs. 5 and 6 show typical spectra of the impedance and pressure of the sound source, respectively, when the injection system is normal, i.e. the injection open pressure is set to 250 bar and the exhaust valve clearance is set to 0.35 mm. To correlate the data to the engine source, the frequency axis is normalised by the engine fire frequency (FF) which is twice the engine speed in hertz for a common fourcylinder diesel engine. Both the real and imaginary parts of the impedance in Fig. 5 show that impedance values have many clear peaks over the frequency range but a basic trend, shown by three slow fluctuations, can be observed from the imaginary part. A further study, conducted by examining the behaviour of these two features over different engine operating conditions, has found that the basic trend has similar behaviour over different engine loads and speeds. However, both the amplitudes and frequencies of the peaks change considerably in a random manner. In particular, the frequency value of the peaks does not have any connection with the FF and its high

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Source Impedance 10 Speed=1507 r/m

Load=44 Nm Opening Pressure=250 bar

Real

5 0 -5 -10 0

1

2

3

4

5 Fire Order

6

7

8

9

10

9

10

10 Speed=1507 r/m

Load=44 Nm Opening Pressure=250 bar

Imaginary

5 0 -5 -10 0

1

2

3

4

5 Fire Order

6

7

8

Fig. 5. Spectrum of source impedance.

Source Pressure 260 250 240

Source Pressure (dB)

230 220 210 200 190 180 170 Speed=1507 r/m Load=44 Nm Opening Presssure=250 bar 160 0

1

2

3

4

5 6 Fire Order

7

8

9

10

Fig. 6. Spectrum of source pressure.

order harmonics. Because the basic trend is insensitive to changes in engine conditions and the peaks have unpredictable behaviours, source impedance may not be a good indicator for engine fault detection. Therefore, the study focuses on the source strength in developing detection features.

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Fig. 6 shows the spectrum of the source pressure. The FF component and its harmonics at the integer order numbers are clearly the dominant components of the spectrum, showing that the spectrum has strong connections with the engine working process. In addition, there are many components at the fractional orders: 1/4, 2/4 and 3/4. This shows that the engine combustion is not uniform over the four cylinders, showing this engine is imperfect working condition. Because of the manufacturing tolerances, the four cylinders can never operate in exactly the same manner. In addition, the exhaust manifold geometry also has some degree of asymmetry. The fractional component is thus inevitable but with acceptable amplitudes for healthy engines. Nevertheless, the amplitudes of the fractional components can be a good indicator for engine health detection, higher amplitudes indicate poorer engine condition. Once the fractional components reveal that some of the cylinders are in a faulty condition, it is required to localise the fault, i.e. to diagnose which of the cylinders are abnormal. For this reason, the results are examined mainly in the time domain using the crank angle. This is because the fluctuations of the waveform are well correlated with individual cylinders and faulty combustion cylinders can be localised easily. Figs. 7–9 show the source pressure waveforms of the engine without any faults. It is seen from these figures that there are four major fluctuations in one engine cycle (0–7201), each corresponding to one of the four cylinders. As

x 104

Pressure(Pa)

2

Residual pressure

Piston motion

1 0 -1 E1O

-2 0

180

E1C 360

540

720

540

720

540

720

x 104

Pressure(Pa)

2 1 0 -1 E1O

-2 0

180

E1C 360

x 104

Pressure(Pa)

2 1 0 -1 E1C

E1O

-2 0

180

360 Crank Angle(Deg)

Fig. 7. Source pressure waveforms for different load conditions at 1500 rpm. (a) Open pressure 250 bar at 1494.7 r/m 22.2 Nm, (b) open pressure 250 bar at 1507.7 r/m 43.2 Nm, (c) open pressure 250 bar at 1504.1 r/m 63.2 Nm.

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x 104

Pressure(Pa)

2 1 0 -1 E1O

-2 0

180

E1C 360

540

720

540

720

540

720

x 104

Pressure(Pa)

2 1 0 -1 E1O

-2 0

180

E1C 360

x 104

Pressure(Pa)

2 1 0 -1 E1C

E1O

-2 0

180

360 Crank Angle(Deg)

Fig. 8. Source pressure waveforms for different load conditions at 1800 rpm. (a) Open pressure 250 bar at 1801 r/m 22.1 Nm, (b) open pressure 250 bar at 1822.6 r/m 41.8 Nm, (c) open pressure 250 bar at 1805.5 r/m 61.8 Nm.

the vertical grid lines represent the TDC (TDC/BDC) of each cylinder, 1801, 3601, 5401 and 7201 (01) are thus in-line with cylinders 1, 2, 4 and 3, respectively, which is also the firing sequence of the engine. More details of the pressure fluctuation can be observed respective to each cylinder. As illustrated by the horizontal solid line in each plot, the exhaust valve of cylinder 1 opens in advance of BDC and closes beyond TDC, referring to cylinder 1. This allows an understanding of the waveform fluctuation feature for a particular cylinder. The two small pressure peaks corresponding to each cylinder are formed by two characteristic exhaust effects, respectively. The first one is the exhaust flow driven by the cylinder residual pressure at the moment of exhaust valve opening and the second is the exhaust flow forced by the upward piston motion. Because the exhaust valve opens in advance of BDC and closes beyond ADC, the exhaust pressure from the first characteristic effect adds to the pressure from the second effect. The first exhaust effect, resulting from residual expansion upon the exhaust valve first opening, causes a burst of pressure in the exhaust pipe and a sharp rise peak in the waveform within around 201 before BDC of crank rotation. This is followed by a steady decrease in pressure as the residual pressure pulse dissipates, before the second exhaust effect associated with the upwards movement of the piston starts to take effect. As a result, the pressure waveform rises to a second peak at around 701 before the TDC.

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x 104

Pressure(Pa)

2 1 0 -1 E1O

-2 0

180

E1C 360

540

720

540

720

540

720

x 104

Pressure(Pa)

2 1 0 -1 E1O

-2 0

180

E1C 360

x 104

Pressure(Pa)

2 1 0 -1 E1C

E1O

-2 0

180

360 Crank Angle(Deg)

Fig. 9. Source pressure waveforms for different load condition at 2100 rpm. (a) Open pressure 250 bar at 2104.3 r/m 22.5 Nm, (b) open pressure 250 bar at 2109.4 r/m 41.7 Nm, (c) open pressure 250 bar at 2100.7 r/m 59.9 Nm.

Furthermore, from the pressure waveform in the figures, a number of features can be found by comparing them over different operating conditions. Firstly, the large four fluctuations in one engine cycle are close to each other. This shows that the combustions in the four cylinders are uniform and hence indicates that the engine runs under healthy condition. Secondly, the amplitude of the waveform portion due to the residual pressure exhibits a gradual increase with the increase in engine load but little with engine speeds. This shows that the source strength correlates with combustion strength. In other words, higher cylinder pressures produce higher amplitudes of the residual pressure and hence higher amplitude of the exhaust pressure. Therefore, the amplitude can be a good indicator of the combustion process. Finally, the wave profile corresponding to each cylinder varies with different operating conditions including both the loads and the speeds. The peak around due to the residual pressure increases with load. In contrast, the waveform portion due to upward piston motion exhibits a slight decreasing trend with increase in engine load. In addition, at high engine speed, the exhaust flow forced by the piston motion upwards is more pronounced and hence the sharp pressure rise due to the residual pressure becomes smaller. These changes show again that the exhaust source pressure is sensitive to changes of engine condition and hence can be relied on for fault detection. In particular, a scheme of evaluation of waveform uniformity is used for abnormal cylinder detection and diagnosis.

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4.2. Monitoring of engine combustion In order to evaluate the performance of this method for engine condition monitoring, an injector fault and an exhaust valve fault are introduced into cylinder 1, respectively, with different opening pressures and different valve clearances while the other cylinders are kept the same. Normally, the opening pressure is set to 250 bar and the exhaust valve clearance is set to 0.35 mm for best fuel atomisation and efficient combustion. These settings can deviate from the above values due to wearing of the device, blockages or leakages in the fuel supply line. These faults will cause reduced engine performance and produce more emissions. In this study, four different opening pressures 150, 190, 250 and 280 bar and three different valve clearance 0.35, 0.7 and 1.7 mm are used to examine the detection sensitivity and diagnosis capability based on the exhaust source waveform.

Pressure(Pa)

x 104 2

0 E1O

E1C

-2 0

180

360

540

720

540

720

540

720

540

720

Pressure(Pa)

x 104 2

0 E1O

-2 0

180

E1C 360

Pressure(Pa)

x 104 2

0 E1C

E1O -2 0

180

360

x 104 Pressure(Pa)

2

0 E1C

E1O -2 0

180

360 Crank Angle(Deg)

Fig. 10. Waveform of exhaust source at different opening pressures: 150, 190, 250 and 280 bar. (a) Injector open pressure 150 bar at 1505.2 r/m 22.4 Nm, (b) injector open pressure 190 bar at 1496.1 r/m 21.4 Nm, (c) injector open pressure 250 bar at 1496.1 r/m 22.4 Nm, (d) injector open pressure 280 bar at 1502.5 r/m 23.7 Nm.

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Fig. 10 shows the waveforms obtained at the four different opening pressures, respectively. Comparing with the waveform of the healthy condition, shown in Fig. 10(c), the three faulty conditions shown in Fig. 10(a), (b) and (d) can be detected easily by checking the uniformity of its own waveform over the engine cycle. It can be seen from the waveforms that the peak value of the faulty cylinder shows higher amplitude because of the higher residual pressure of retarded combustion. Based on this, a primary detection and diagnosis feature can be developed by evaluating the degree of deviation of the waveform portion between 1801 and 3601. In general, larger deviations correspond to more severe injector faults. As shown in the figures, the largest deviation is with the injector opening pressure of 150 bar, which is 100 bar lower than the required 250 bar. The medium deviation is from 190 bar injector which is 60 bar lower, while the least deviation is from the 280 bar injector which is 30 bar higher. In addition, the angular position of the deviated waveform portion is between 1801 and 3601 and is just in-line with the exhaust stroke of cylinder 1. It is obvious then that the fault must be in cylinder 1. Fig. 11 shows the exhaust waveform at high engine load and speed for the faulty injection pressure of 190 bar. The deviation of the waveform portion between 1801 and 3601 from rest three is clear, demonstrating that this method can be used to detect and diagnose this medium fault in a wide range of operating conditions. However, it has found that the method cannot reveal the deviation from the small injector fault of 280 bar open pressure at engine high load and high speed. This is because the combustion is close to each other when the fuel amount is high under the high engine output condition. On the other hand, this means this method is convenient and easy to use because it does not need high load applied to the engine, which is often difficult to achieve, for the detection of faults. Figs. 12 and 13 show the waveforms of the exhaust source when the exhaust valve clearance is lager than the normal condition, 0.35 mm. As shown in both figures, it is obvious that the waveform of cylinder 1 is different from the others, because the valve of cylinder 1 opens later and closes earlier than others, and thus it exhausts less than it should. This effect is more pronounced when the valve clearance is 1.7 mm. As shown in Fig. 13, the waveform portion corresponding to upward piston motion has a significant drop.

x 104

Pressure(Pa)

2 1 0 -1 E1O

-2 0

180

E1C 360

540

720

540

720

x 104

Pressure(Pa)

2 1 0 -1 E1O

-2 0

180

E1C 360 Crank Angle(Deg)

Fig. 11. Waveforms of exhaust source at high load and high speed at faulty opening pressure of 190 bar. (a) Injector open pressure 190 bar at 1493.4 r/m 62.5 Nm, (b) injector open pressure 190 bar at 2097.7 r/m 64.3 Nm.

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Fig. 12. Waveforms of exhaust source for different loads at faulty valve clearance 0.7 mm. (a) Exhaust valve clearance 0.7 mm at 1505.7 r/m 20.9 Nm. (b) exhaust valve clearance 0.7 mm at 1499.4 r/m 39.1 Nm, (c) exhaust valve clearance 0.7 mm at 1503.4 r/m 63.8 Nm.

Compared with the waveform deviations from fuel injector faults in Figs. 10 and 11, the waveform drop is more distinctive for the change of valve clearance. Therefore, this drop feature can be relied to differentiate the two types of faults investigated.

5. Conclusion In this paper, an effective approach to engine condition monitoring is developed based on the concept of acoustic one-port source. A two-load procedure is studied both theoretically and experimentally for a diesel engine exhaust source measurement. This method minimises the influence of wave reflection from the exhaust system and produces a more accurate description of the engine source. Experimental results show that the exhaust source impedance is not as sensitive to engine operating conditions as is its source strength (pressure). The source pressure is then taken as the only signal for condition monitoring information extraction. Both the spectrum and the waveform can be used for fault detection but the latter allows the extraction of direct diagnoses information regarding the fault location and severity. The experimental results show that two common faults: reduced injection pressure and increased valve clearance can be accurately detected, localised and differentiated by using the local deviations of the waveforms. It has then been demonstrated that a one-port acoustic source-based measurement and analysis is an effective approach to the combustion

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Fig. 13. Waveforms of exhaust source for different loads at faulty valve clearance 1.7 mm. (a) Exhaust valve clearance 1.7 mm at 1505.5 r/m 22.4 Nm, (b) exhaust valve clearance 1.7 mm at 1501.9 r/m 39.6 Nm, (c) exhaust valve clearance 1.7 mm at 1505.8 r/m 65.9 Nm.

characterisation and hence provides a novel and yet robust method of fault detection and diagnosis for the condition monitoring of diesel engines.

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