combustion timing determination based on vibration velocity in hcci engines
TRANSCRIPT
Combustion timing determination based on vibration velocity inHCCI engines
Yong Cheng a,⁎, Juan Tang a,b, Shaobo Ji a, Minli Huang a
a College of Energy and Power Engineering, Shandong University, Jinan 250061, PR Chinab School of Automobile and Transportation Engineering, Liaocheng University, Liaocheng 252004,PR China
a r t i c l e i n f o a b s t r a c t
Article history:Received 7 October 2011Received in revised form 21 June 2012Accepted 1 August 2012Available online 28 August 2012
Combustion timing is one of the most important feedback parameters for combustion controlof homogeneous charge compression ignition (HCCI) engines. It is well known that the var-iation of combustion affects the engine outer surface vibration. Hence, the present work aimsat investigating the possibility of using cylinder head's vibration response as a means todetermine the combustion timing of HCCI engines. This paper proposes a new combustiontiming identification method based on measured vibration velocity. In order to provide aclearer insight into the relationship between cylinder head's vibration velocity and the com-bustion timing, a finite element model (FEM) of a 295 HCCI engine is established. Based on theFEM simulation, the vibration velocity excited by combustion is simulated and results showthat the cylinder head's vibration velocity and the rate of in-cylinder pressure rises havesimilar trend before the peak pressure occurs. The turning point of the cylinder head vibrationvelocity curve during the compression stroke corresponds to the combustion timing. Ex-periments are conducted on the 295 HCCI engine to compare and verify with the FEM results.Test results confirm that when the combustion begins, a turning point will occur in thecylinder head's vibration velocity curve. Therefore the combustion timings can be detected bythe turning point of the cylinder head's vibration velocity during the compression stroke.Comparison is made between the combustion timings that are determined from the vibrationvelocity and from the rate of cylinder pressure rise. Compared to the combustion timingsdetermined from the rate of cylinder pressure rise, the accuracy of that from the turning pointof the cylinder head vibration velocity is within ±0.5°CA.
© 2012 Elsevier Ltd. All rights reserved.
Keywords:HCCI engineCylinder head vibration velocityCombustion timing determination
1. Introduction
In recent years, HCCI engines have drawn great attention as its obvious advantages over conventional compression ignition(CI) or spark ignition (SI) engines. HCCI engines typically have high compression ratios so that it can produce thermal efficienciescomparable to diesel engines. It could also operate at much leaner mixtures which results in much lower nitrogen oxide andparticulate emissions than conventional engines [1–3]. Besides these benefits, however, some operational issues arise, especiallyfor the control of the ignition and combustion. In a HCCI engine, ignition occurs at multiple locations throughout the cylindervolume, starting at the hottest regions. As the mixture in the hottest region starts to burn, the expanding gasses cause furthercompression of the un-burnt mixture and subsequent ignition occurs without flame propagation [4]. Therefore the combustion ismainly controlled by chemical kinetics and has no external control. For control of combustion in the cycles to follow, the currentignition timing of a HCCI engine must be sensed first.
Mechanism and Machine Theory 58 (2012) 20–28
⁎ Corresponding author. Tel.: +86 531 8839 6530.E-mail addresses: [email protected] (Y. Cheng), [email protected] (J. Tang), [email protected] (S. Ji), [email protected] (M. Huang).
0094-114X/$ – see front matter © 2012 Elsevier Ltd. All rights reserved.http://dx.doi.org/10.1016/j.mechmachtheory.2012.08.004
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Combustion timing can be estimated accurately from the in-cylinder pressure traces which can be directly measured bypiezoelectric pressure transducers [5,6]. Unfortunately, high precision piezoelectric pressure transducers are usually expensiveand not feasible for all situations, especially for on-line applications. Compared to dynamic pressure transducers, vibration sensorsor ion sensors are significantly less expensive andmore adaptive for on-line applications. For a HCCI engine that is modified from aSI engine, methods based on ion signals may be more suitable. Mehresh et al. [7] developed a numerical model to analyze thepossibility of identifying combustion timing based on ion signals. Results show that there is a correlation between the cylinderpressure rise rate and the ion signal. The ion signal is a suitable measure of combustion timing and could be used for feedbackcontrol of HCCI engines. However, for a HCCI enginemodified from a CI engine, methods based on ion signals are not as convenientasmethods based on vibration or acoustic signals. Gu et al. [8,9] simulated the sound level of a diesel engine during the combustionprocess, and utilized time to frequency domain analysis to reveal the underlying characteristics of the sound waves. Resultsshowed that the combustion events can be observed from the sound signal above 20 kHz. Grondin et al. [10] proposed amethod toestimate the combustion timing of a Diesel HCCI engine for controlling purpose. This combustion status estimator works as a highfrequency wavelet identifier that takes as input of the vibrating acceleration signal recorded on the engine block with a knocksensor. The authors defined a band pass filter to separate the combustion from other noise sources on the vibrating accelerationsignal. The low and high cut-off frequencies of the band pass filter are set as 1000 Hz and 2500 Hz. The medium point of thecombustion process was related to the position where the instantaneous energy of the filtered vibration signal is attainedmaximum. The estimated location of themiddle of combustion is defined as the center point of the Gaussian function. Arnone et al.[11] showed that regardless of the engine's load conditions, high values of coherence are exhibited between the vibration signals inthe vertical direction with the in-cylinder pressure in the frequency band of approximately 500 Hz–1100 Hz. When the initialignition of air–fuel mixture takes place, the in-cylinder process are characterized by occurrences of frequency components of high
Fig. 1. Finite element model of the 295 engine.
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amplitudes in the 500 Hz–1000 Hz range at some degrees before TDC. Such a behaviormakes it possible to identify the combustionevents using vibrating acceleration signals. However, methods based on vibrating acceleration signals require to analyze thecorresponding frequency band with the combustion, different researchers have different views. In addition to those methodsdiscussed above, some researchers predict combustion events based on mathematical models with easily measurable inputs, andapplied for various engine operating conditions. For instance Shahbakhti et al. [12] applied a modified knock integral model(MKIM)which contains the correlated gas exchange process and heat releasemodels, to predict HCCI engines' combustion timing.The MKIMmodel is parameterized with a thermal-kinetic simulation model. Results show that this model is suitable for real timeimplementation and is able to predict ignition timing with an average error of less than 2 crank angle degrees. Model basedmethods which take into account of engine operating conditions such as engine speed and EGR ratio, etc. are more complex thanthe methods based on vibration signals or ion signals.
Presented in this paper is a new method based on processing of vibration velocity signals. Firstly discussed is the possibility ofutilizing vibration velocity to identify the combustion timing. The relationship between the in-cylinder pressure traces and thevibration velocity signals is then demonstrated. Themethod is applied on a 295 HCCI engine by utilizing the FEM simulated vibrationvelocity to identify the combustion timing. Experiments are then conducted on the same engine to validate this new method.
2. Finite element model
In internal combustion engines, themeasured vibration velocity of the cylinder head is a response to the combined excitation ofthe combustion, piston slap, valve impact and other sources. Identification of the combustion events directly from the measuredvibration velocity is usually difficult. However, numerical method can separate the vibration velocity excited by the combustionalone. One can easily characterize the correlation between the vibration velocity and the combustion event. The numerical methodsuch as finite element method (FEM) has been widely used to determine the dynamic response of a vibrating structure [13–16]. Inthis paper the ABAQUS software is applied.
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Fig. 2. In-cylinder pressure and vibration velocity of the three measuring points on cylinder head.
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Fig. 3. Amplitude response of the 295 HCCI engine and the spectrum of the in-cylinder pressure.
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To achieve a reliable prediction of the vibration excited solely by the combustion, a comprehensive FEM model is requiredwhich takes into account of nearly all the construction details. However, the process of modeling in this way would be verytedious and the required computational time is also a problem. Owing to the fact that our major concern here is the cylinder headvibration induced by combustion, the FEM model only needs to include the engine block (HT250) and cylinder head (QT300), asthese parts are the dominant contributors to the combustion induced vibration. The FEM model is shown in Fig. 1.
The engine is “mounted” on the test bench by four supporters. In simulation the supporters are assumed rigid bodies and haveno deformations in all DOFs. Thus the supporters are constrained in all DOFs, as shown in Fig. 1. To eliminate the interference ofpiston-slap, the main bearing reaction forces, the valve impact, as well as other no-combustion related excitations, only the in-cylinder pressure is taken into account. The measured in-cylinder pressure trace is applied on the combustion chamber surface. Inthis way the vibration responses induced by the change of cylinder pressure can be simulated.
In addition, the model is developed with some other assumptions. The material property is assumed ideally elastic. This is areasonable assumption for steel block and cylinder head since the stress level exhibited is well below the yielding limits. Also thisassumption has been previously proven feasible in other references [17,18]. It is also assumed that the damping ratio in thedampingmodel remains constant throughout the simulation. This assumption is mainly for simplification of the FEM solver, whichis typically the case for most vibration problems.
The finite element equations of motion, which represent the structural vibration in response to internal combustion pressurepulsations, can be expressed as:
M½ � €qf g þ D½ � _qf g þ K½ � qf g ¼ R _q; q; tð Þf g ð1Þ
where [M], [D] and [K] are the mass, the damping and the stiffness matrices of the engine;
€qf g, _qf g and {q} are the time-dependent displacement, speed and acceleration responses;R _q; q; tð Þf g is the time-dependent external force vector, here means the in-cylinder pressure which is measured by the
piezoelectric pressure transducer in the 295 HCCI engine.
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Fig. 4. Vibration of cylinder head and in-cylinder pressure.
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Fig. 5. The vibrating speed and the second derivative of the rate of pressure rise under different conditions.
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This displacement-based FEM formulation has previously proven to work with sufficient accuracy for this type of simulation[18]. For a dynamic analysis Eq. (1) is solved using a direct time integration method. In this simulation the algorithm is employingimplicit methods, such as the Newton integral method in the ABAQUS software.
3. Numerical results and discussion
3.1. Vibration responses of different output nodes
For the purpose of identifying the most sensitive vibration points to the combustion process, three vibration output nodes arechosen on the cylinder head, named the first, the second and the third points, as shown in Fig. 1.
The operating condition of the 295 HCCI engine is controlled by the fuel quantity of two consecutive injections. The operatingcondition is expressed by the speed and the two consecutive injection pulse widths. For instance, 1400–300–700 represents anoperating state that has an engine speed of 1400 r/min, with the pilot injection duration of 300 μs and the secondary fuelinjection duration of 700 μs. Setting the first cylinder's firing top dead center (TDC) as 0 degree crank angle, in Fig. 2 the vibrationvelocity simulated for the three output nodes on the cylinder head under the 1400–300–2000 operating condition are comparedwithin the crank angle from−60°CA to 60°CA. It can be easily concluded that the vibration velocity of the three nodes excited bythe first cylinder's combustion have similar tendency but different amplitudes. The amplitude of the first node's vibration velocityis the largest compared to the other two. This is because the first node is the closest to the first cylinder and the energy loss is thesmallest. Thus the first node is chosen to analyze cylinder #1's combustion process since it has the most sensitive vibration.
3.2. Relationship between cylinder head vibration velocity and the in-cylinder pressure profile
According to the Hooke's Law, the deformation of any structure, if ideally elastic, is proportional to the force acting on it. If thestructure is ideally rigid, according to Newton's law the acceleration is proportional to the force acting on it. In reality no structureis ideally elastic nor ideally rigid, but something in between and the relationship between the vibration displacement and theexciting force is determined by the amplitude–frequency characteristics and the spectrum of the excitation force.
Table 1Start of combustion determined from the simulated vibration velocity and the rate of pressure rise (°CA).
Operating conditions Start of combustion determined fromthe rate of pressure rise
Start of combustion determined fromthe vibration velocity
Deviation
1400–300–2000 −5.8 −3.2 2.61400–300–1700 −4.7 −2.5 2.21400–300–1300 −5.7 −3.4 2.31400–300–900 −13.6 −11.2 2.41400–300–700 −11.8 −8.2 3.6
Note: the firing TDC is designated as 0°CA.
Table 2295 HCCI engine's characteristics.
Engine type In-line, 4-stroke naturally aspirated
Cylinder TwoBore (mm) 95Stroke (mm) 115Compression ratio 13:1Injector High-pressure swirl-type nozzleInjection type Direct injectionInjection pressure (MPa) 6.0Fuel injection rate (mm3/ms) 15.7Spark plug type A7RCIgnition energy (mJ) 70
Table 3Experimental transducers and associated technical data.
Test devices Type Characteristics
Pressure transducer 12QP250 Piezoelectric, sensitivity: 200 pC/MPa, measuring range: 15 MPaCharge amplifier YE5850A Extremely low frequency: 2 μHz, gain: 0.1–1000 mV/Unit, sensitivity: 1–109.9 pC/unitVibration velocity transducer CS-YD-005 Piezoelectric, measuring range: 200 mm/s, measuring frequency: 4 Hz–4 kHz, sensitivity: 50 mV/mm/sData acquisition system USB2002 16 channels,400 kHz A/D card, measuring range: ±10 V, measuring frequency: 50 kHz each channel
24 Y. Cheng et al. / Mechanism and Machine Theory 58 (2012) 20–28
The 295 HCCI engine's amplitude–frequency characteristics can be obtained by the FEM model while the spectrum of thein-cylinder pressure can be obtained by Fourier transformation. The amplitude–frequency characteristics of the 295 HCCI engineand the spectrum of the in-cylinder pressure are shown in Fig. 3(a) and (b) respectively. It can be seen from Fig. 3(b) that thein-cylinder pressure is mainly composed by the harmonic terms below 650 Hz. The FFT amplitude of the harmonic terms beyond650 Hz is smaller than 0.013. According to the amplitude–frequency characteristic curve, when the exciting force frequency isbelow 650 Hz, the vibration displacement amplitude ratio varies from 0.8 to 2.1. This value varies in a small range and is veryclose to 1. So the vibration displacement may nearly be proportional to the in-cylinder pressure or the tendency of vibrationdisplacement and the in-cylinder pressure may be similar.
Shown in Fig. 4 are the first node's vibration displacement and velocity simulated by the FEMmodel, the measured in-cylinderpressure trace, as well as the rate of the pressure rise under the1400–300–2000 operating condition. The tendency of thevibration displacement is similar to the tendency of the in-cylinder pressure which is coincided with the above analysis. For thevibration velocity is the derivative of the vibration displacement, there is a similar tendency between the vibration velocity andthe rate of pressure rise. Also there is a phase difference between the vibration displacement to the in-cylinder pressure, and thevibration velocity to the rate of pressure rise. The value of the phase shift is determined by the phase–frequency characteristics of
Fig. 6. Experimental apparatus.
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Fig. 7. In-cylinder pressure, the rate of pressure rise and the vibrating speed under different operating conditions.
25Y. Cheng et al. / Mechanism and Machine Theory 58 (2012) 20–28
the vibration system and the spectrum of the in-cylinder pressure. For a specific engine, the phase–frequency characteristic isconstant and the spectrum of the in-cylinder pressure varies only in a small range. Thus the phase shift can be considered assystem deviation.
3.3. Combustion timing identification based on the vibration velocity
The start of combustion can be identified from the rate of pressure rise. The turning point in the rate of pressure rise during thecompression stroke can be considered as the start of combustion [19]. Since the vibration velocity and the rate of pressure risehave similar tendency when combustion starts, the turning point in vibration velocity can also be treated as the combustiontiming. Thus the turning point of the vibration velocity in the compression stroke can be used to identify the combustion timing.Fig. 5 shows the simulated vibration velocity and the rate of the pressure rises under different operating conditions. At thelocation where the combustion starts, an obvious turning point can be seen in the vibration velocity and the rate of pressure risecurves. The turning points are marked as φig to present the position where combustion starts. Table 1 shows the combustion startpositions under different conditions determined from the simulated speed and the rate of pressure rise. From Table 1 we can seethat there is a phase shift between the combustion timings determined from the vibration velocity and that from the rate ofpressure rise. For the 295 HCCI engine, with the change of operating conditions, the phase shift varies from 2.2°CA to 3.6°CA.Considering the averaged value of 2.9°CA as the system deviation and taking the start of combustion identified from the rate ofpressure rise as the reference, the error on the start of combustion identified from the vibration velocity is less than ±0.7°CA. Thisindicates that it is feasible to identify the combustion timing from the simulated vibration velocity.
4. Experimental verification and discussion
To validate the simulation results, experiments are conducted on a 295 HCCI engine, which is refitted from a two-cylinderdiesel engine. Table 2 shows the main characteristics about the 295 HCCI engine. The engine is installed on a test bench andinstrumented for indicated signal measurements like in-cylinder pressure and vibration velocity signals. Table 3 lists out theexperimental transducers and the corresponding technical specifications. Fig. 6 shows the test bench setup.
Fig. 8. Time–frequency analysis of the vibrating speed under different operating conditions.
26 Y. Cheng et al. / Mechanism and Machine Theory 58 (2012) 20–28
The in-cylinder pressure traces and the vibration velocity at the first node are measured on cylinder #1. Fig. 7 depicts thein-cylinder pressure traces, the rate of pressure rise and the vibration velocity under different operating conditions. Experimentaldata show that, except for the motoring condition, the rate of pressure rise and the vibration velocity have similar tendency whenthe combustion starts, which verifies the simulation results. However the influences of the disturbance signal cannot be ignored,especially at the position of 45°CA before TDC. In order to extract out the vibration velocity excited by the combustion, S transformmethod is applied to analyze the time–frequency characteristic of the vibration velocity, as shown in Fig. 8. It can be seen thatwhen the combustion starts, the vibration velocity is mainly concentrated in the band below 2000 Hz, while the vibration velocityis mainly concentrated in the band over 2000 Hz when the disturbance signal occurs. It can be concluded that the vibrationvelocity below 2000 Hz is highly correlated to the combustion process. Thus a low-pass digital filter is designed to extract out thesignals below 2000 Hz. The filtered vibration velocity signals are shown in Fig. 9.
The combustion timing can be identified from the filtered vibration velocity. Table 4 depicts the combustion timings determinedfrom the vibration velocity and the rate of pressure rise. The combustion timing determined from the vibration velocity has alsoshown a delay in the phasing, compared to the combustion timings identified from the rate of pressure rise. The phase delay variesfrom 2.4°CA to 3.3°CA and the averaged value is 2.85°CA, the latter can be treated as the system deviation. Taking the rate of pressurerise method as the datum, the combustion timing determination precision from the vibration velocity is reduced to ±0.45°CA.Experimental results also indicate that combustion timing can be effectively determined from the cylinder head vibration velocity.
5. Conclusions
Through finite element modeling, the relationship is revealed for a HCCI engine between the cylinder head's vibration velocityand the start of combustion. The method is presented and the major conclusions can be summarized as follows:
1) The vibration velocity has a similar tendency to the rate of the cylinder pressure rise. The turning point of the vibrationvelocity in the compression stroke can be treated as the start of combustion.
2) There is a phase shift between the vibration velocity and the rate of cylinder pressure rise. The phase differences at differentengine operating conditions only vary in a small range. The averaged value can be considered as a system deviation.
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Fig. 9. Cylinder #1: cylinder pressure, rate of pressure rise and the vibration velocity at different operating conditions.
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3) For the 295 HCCI engine, taking the rate of pressure rise as the reference, the start of combustion determination error by thevibration velocity is limited to ±0.5°CA.
Acknowledgments
This work was supported by National Natural Science Foundation of China (NSFC: 50876055) and the support is gratefullyacknowledged.
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Table 4Start of combustion determined from vibration velocity and the rate of pressure rise (°CA).
Conditions Determined from the rate of pressure rise Determined from vibration velocity Deviation
1200–300–1500 −0.9 1.8 2.71400–300–700 −6.0 −2.7 3.31400–300–900 −6.5 −3.5 3.01400–300–1100 −2.9 −0.2 2.71400–300–1300 −2.7 −0.3 2.41400–300–1500 −2.7 −0.3 3.01400–300–1700 −2.1 0.6 2.71400–300–2000 −3.6 −0.9 2.71600–300–1500 −5.1 −2.4 2.7
28 Y. Cheng et al. / Mechanism and Machine Theory 58 (2012) 20–28