Cylinder Pressure-Based Combustion Controls for Advanced Diesel Combustion With Multiple-Pulse Fuel Injection

IEEE TRANSACTIONS ON CONTROL SYSTEMS TECHNOLOGY, VOL. 21, NO. 6, NOVEMBER 2013 2143

Cylinder Pressure-Based Combustion Controls forAdvanced Diesel Combustion With

Multiple-Pulse Fuel InjectionIbrahim Haskara and Yue-Yun Wang

Abstract— For current diesel engines, multiple fuel injectionmechanisms enabled by high rail-pressure systems is a key leverthat can help to achieve further reduction in engine-out emissionsand improvements in performance. In the case of multiplefuel injections, timing and fuel pulse-width for each pulse (or,equivalently, fuel amount) need to be optimized and maintainedfor low emissions, fuel economy, noise, and exhaust thermalmanagement over different operating ranges. This paper presentsa research study on the application of pressure-based controlsfor management of the multiple-pulse fuel injection, particularlymain and post injections, to maintain a robust combustionbehavior against disturbances and variations in the field. Severalcontrol features for simultaneous management of main and postinjections are proposed and experimentally validated on a 6.6 LV8 diesel engine in an engine dynamometer, both at steady-stateand during federal test procedure transients.

Index Terms— Automotive systems, closed-loop combustioncontrols, engine controls.

I. INTRODUCTION

AMAJOR challenge being faced in diesel technology ismeeting current and future emission requirements with-

out compromising fuel economy. Unlike standard gasolineengines, in which a three-way catalyst with lambda-control(air–fuel ratio control) is the well-known and mature approachto meet the emission regulations, the diesel emission challengeis an open research area that is being investigated from differ-ent perspectives. Particularly, emissions of oxides of nitrogen(commonly known as NOx) are a main concern of lean-burndiesel engines. Lean NOx traps and selective reduction catalystsystems are two NOx aftertreatment solutions. Similarly, par-ticulate matter and unburned hydrocarbon emissions also callfor specific aftertreatment devices such as diesel oxidationscatalysts and diesel particulate filters.

Recently, several low-temperature combustion concepts[homogenous charge compression ignition (HCCI), premixedcharge compression ignition (PCCI), etc.] have also beeninvestigated to reduce engine-out emissions, which wouldreduce the burden of a costly aftertreatment system. Thebasic characteristic of these special combustion modes is

Manuscript received December 8, 2011; revised August 8, 2012; acceptedOctober 6, 2012. Manuscript received in final form October 26, 2012. Dateof publication December 11, 2012; date of current version October 15, 2013.Recommended by Associate Editor J. Lu.

The authors are with GM Global Research and Development, Warren, MI48090 USA (e-mail: [email protected]; [email protected]).

Color versions of one or more of the figures in this paper are availableonline at http://ieeexplore.ieee.org.

Digital Object Identifier 10.1109/TCST.2012.2227057

to reduce the bulk cylinder gas temperatures by using veryhigh exhaust gas recirculation (EGR) levels and also relyingon multiple-pulse fuel injection strategies to create a morehomogenous mixture in the combustion chamber before theonset of the combustion. For an EGR system, a portion ofalready burned inert exhaust gases is either recirculated backto the intake manifold (external EGR) or trapped inside thecylinder via engine valve timing and lift (internal EGR) so asto mix with the fresh air charge of the next combustion cycle.Dilution increases the thermal capacity of combustion mixtureand lowers the peak combustion temperature which, in turn,reduces the NOx formation rate during combustion. Further-more, it enables fully premixed combustion by prolonging theignition delay beyond the fuel injection duration, for whichthe combustion initiates solely based on the thermodynamicconditions in the cylinder. In addition, multiple-pulse fuelinjections bring additional combustion optimization capabilityfor further reduction in engine-out emissions through betterheat release shaping during a combustion event. Althoughsignificant engine-out emission reduction is achievable, thesensitivity of these combustion modes requires a precisecontrol of mixture and fuel injection profiles to make theseadvanced combustion engines viable in production.

Combustion feedback controls have been investigated forboth gasoline and diesel engine applications in the past; how-ever, their commercial applications were very limited beforethe advent of advanced combustion modes (HCCI, PCCI),in which the combustion feedback controls appear as a keyenabler. Closed-loop (CL) spark timing control for maximumbrake torque (MBT) can be considered the earliest gasolineengine application, where a pressure sensor was used to obtainempirical MBT timing indicators. Locating peak pressurearound 15° after top dead center (TDC), achieving 50% massfraction burned at 8°–10° after TDC, or controlling normalizedpressure ratio of in-cylinder and motoring pressures at 10°after TDC [PR(10)] around 0.55 to obtain the MBT timing areexamples of such empirical criteria reported in the literature[1]–[8]. Among the most recent diesel engine applications,CL control of injection timing and EGR amounts on thebasis of cylinder pressure feedback to enable advanced dieselcombustion (such as PCCI) has also been reported. To thisobjective, pressure signal has been processed in real timeto extract certain metrics related to combustion phasing andload [9]–[16]. However, the literature for diesel applicationsis mainly concentrated on the use of cycle-average metrics,

1063-6536 © 2012 IEEE

2144 IEEE TRANSACTIONS ON CONTROL SYSTEMS TECHNOLOGY, VOL. 21, NO. 6, NOVEMBER 2013

such as 50% mass fraction burned (CA50) and indicatedmean effective pressure (IMEP), to adjust main injectiontiming and quantity without any emphasis on the collaborativemanagement of multiple-pulse fuel injection profiles. In caseof multiple injections, timing and pulse-width (fuel amount)of each injection pulse need to be optimized for emissions,noise, and fuel economy. The control system is then expectedto maintain this optimized multistage combustion againstcomponent and fuel variations, disturbances, and aging.

This paper presents several methods in which the cylinderpressure signal is used for multiple-pulse fuel injectionmanagement for a diesel engine capable of running in low-temperature combustion modes. Among multiple-pulse fuelinjection profiles, we concentrate primarily on main and postinjections since fuel amounts injected for the two have thelargest compounding effect on the combustion metrics (CA50and IMEP). A 6.6 L, V-8 diesel engine was used for controlsystem mechanization and experiments. A custom-madecylinder pressure processing unit (CPPU) is developed andincorporated into the engine dynamometer setup. Pressuretransducers are installed in each cylinder’s glow-plug hole andfed through charge amplification module to the CPPU. Theencoder signal is also processed by the CPPU unit to enablecrank synchronous sampling with one crank-angle degreesampling rate. Event-based cylinder processing algorithms areused to compute combustion feedback variables, which areused to adjust the injection timings and pulse-widths for thenext combustion event. Control algorithms are implementedusing rapid prototyping hardware in bypass mode to theproduction engine control module (ECM). The controller iscapable of separately adjusting individual cylinder injectiontiming and fueling quantity at each combustion firing. Thebypass values are used as an addition to the base timing andfueling maps that were available in the ECM calibration for theengine.

II. CONTROL-ORIENTED HEAT RELEASE MODELS

Combustion characteristics can be inferred from cylinderpressure. Extracted heat release information from cylinderpressure is the basis for the pressure-based combustion feed-back control. A heat release curve indicates how a partic-ular combustion event starts and progresses with respect tothe piston position. The heat release generation is the mostdirect process of interpreting the combustion, which producesthe useful torque in addition to the combustion products(emissions). The performance of a combustion event in termsof torque generation, fuel consumption, resulting emissions,noise, and so on can be directly explained by its heat releasecurve. Conversely, all the combustion objectives are realizedby placing the heat release curve as desired, both in terms of itslocation with respect to the piston (phasing) and its magnitude(strength). This is the main motivation of using heat releasefor combustion control purposes.

Fig. 1 shows a basic representation of how cylinder pres-sure can be used to compute heat release and fuel burnedprofiles. The first law of thermodynamics and the ideal gaslaw relate cylinder pressure trace to the corresponding heat

Q

LHVQ1 fuelmnetfuelm ,

Net heat release Energy to fuel mass scaling

∑nv dQPdVdTmc =+mRTPV =

nQP

∫⋅ θddQ

Qht

LHV

1

htfuelm ,f ,

Heat loss and its fuel quantity equivalent

Fig. 1. Schematic of heat release computation from cylinder pressure. Netfuel burned is the scaled heat released.

release profile. Note that net fuel burned is the scaled heatrelease as shown in Fig. 1. Within this basic architecture, thereare different ways of heat release computation, which will bebriefly reviewed next.

A. Traditional Heat Release Analysis

Basic relations for heat release analysis excluding heat lossare given below

mcvdT + PdV = d Qch

PV = m RT

R = cv (γ − 1)

d Qch = QLHVdm f (1)

where P , T , and V are the cylinder pressure, temperature, andvolume respectively, m is the total cylinder mass, Qch is thecumulative heat released, R is the ideal gas constant, cv is theconstant volume specific heat, γ is the ratio of specific heats,QLHV is the lower heating value of the fuel, and m f is theburned fuel mass. Putting all equations together and cancelingout temperature variable with the ideal gas law relation givesthe traditional heat release rate equation [17]

γ

γ − 1· P · dV + 1

γ − 1· V · d P = d Qch. (2)

Heat release is obtained by integrating the above equation.Usually, property dependence of gamma (γ ) is also neglectedfor simplicity by assuming γ a constant∫ (

γ

γ − 1· P · dV + 1

γ − 1· V · d P

)= Qch. (3)

B. Iterative Derivation Including Property Effects

Original heat release equations can be solved in two stepsfollowing the Rassweiler and Withrow approach.

1) Isentropic expansion/compression

mcvdT + PdV = 0. (4)

2) Constant volume heat addition

mcvdT = QLHVdm f . (5)

Fig. 2 shows the two sequentially modeled processes takingplace between the consecutive crank-angle samples of the

HASKARA AND WANG: CYLINDER PRESSURE-BASED COMBUSTION CONTROLS FOR ADVANCED DIESEL COMBUSTION 2145

),,( 111 +++ kkk VTP )( 111 +++ kkk

Isentropic expansion/compression

Constant volume heat addition

Crank angleθ θ

),,( kkk VTP

Isentropic expansion/compression),,( 1expexp +kVTP

Crank anglekθ 1+kθ

Fig. 2. Description of the pressure evolution process between the twoconsecutive crank-angle pressure sampling points.

pressure signal. In Fig. 2, Pk , Tk , and Vk are the cylinderpressure, temperature, and volume at the current crank-angle,respectively, and Pk+1, Tk+1, and Vk+1 are the same variablesat the next crank-angle. Pexp and Texp are the fictitious pressureand temperatures, respectively, that the cylinder gas wouldhave, if the change in the variables between the two samplesis due to only the piston motion (volume change: compressionand expansion).

Solution of isentropic expansion and compression equation(capturing only volume changes) gives

Pexp = Pk ·(

Vk

Vk+1

)γk

(6)

Texp = Tk .

(Vk

Vk+1

)γk−1 ∼= Tk ·(

1 + R

cv· Vk+1

Vk

). (7)

Solution of constant volume heat addition gives

m · cv,Tk+1 · Tk+1 − m · cv,Texp · Texp = QLHV�m f

m ·(

cv,Tk+1 · Pk+1 ·Vk+1

m ·RTk+1

− cv,Texp ·Pexp ·Vk+1

m RTexp

)= QLHV�m f

(1

γTk+1 − 1·Pk+1 − 1

γTexp − 1·Pexp

)·Vk+1 = QLHV�m f

(8)

where the indices in the cv and γ capture the temperaturedependence of these variables. Combining the solutions fromeach process provides the fuel burned (or scaled heat released)between the consecutive crank-angle sampling points

�m f = Vk+1

QLHV

{1

(γTk+1 − 1)· Pk+1

− 1

(γTexp − 1)· Pk ·

(Vk

Vk+1

)γTk

}. (9)

With the final equation, the direct pressure differentiation isavoided, which is suitable for pressure data sampled with low-resolution crank-angles (like 6° as opposed to 1° sampling).The structure also allows the use of variable properties overthe entire range in terms of different operating points runningwith various air and EGR levels as well as variations duringa single combustion event. Different fuel types can also becaptured in γ relation (petroleum diesel, biodiesel, gasoline,ethanol, and so on). The intermediate temperature variablesduring the derivation allow computing these variable mixtureproperties through prescheduled property functions.

200 400 600 800 1000 1200 1400 1600 1800 20004.5

5

5.5

6

6.5

7

7.5

8

Temperature (K)

Spe

cific

heat

(cal

/mol

-K)

Thermodynamic properties (cv)

AirPhi=1, EGR=60% (Pre-combustion)Phi=1, EGR=60% (Post-combustion)Phi=0.2, EGR=60% (Pre-combustion)Phi=0.2, EGR=60% (Post-combustion)

10015.01)(),,( ,,

kkairvkkkv

EGRTcEGRTc

Fig. 3. Specific heat computed at different temperatures and mixturecombinations (air-fuel ratio and EGR percentages). *: an example constantcv value used during the entire combustion.

C. Property Relations

As discussed before, combustion mixture properties arecaptured in the ratio of specific heats, γ , or the specificheat, cv . Specifically, cv quantifies the heat of mixture thatrelates the gas pressure (and temperature) evolution, which ismeasured, to the heat added or extracted during a combustionevent. It varies over combustion because of temperature andmixture changes (air, EGR). Among those, mean gas tem-perature follows pressure changes during combustion, whichcan be computed given an initial temperature prior to com-pression such as intake manifold temperature at intake valveclosing. On the other hand, the closed volume in-cylinderEGR level goes from an operating point-dependent initial EGRlevel to a final amount (100% with complete combustion)in relation to the fuel burning. The crank-angle-resolvedspecific heat of the combustion mixture can be represented asfollows:cv,k(T, φ, rk) = [cv,st(T ) · φ + cv,air(T ) · (1 − φ)] · rk

+cv,air(T ) · (1 − rk) (10)

where cv,air and cv,st are the specific heat capacities for theair and stoichiometric mixtures, which are available fromthermodynamic property tables as a function of temperatureonly, and φ and r are the equivalence ratio and the burnedgas fraction of the mixture, respectively. Equivalence ratioand EGR weight the air and stoichiometric specific heatcapacities to reflect the specific heat for the particular mixture.Figs. 3 and 4 exemplify how the specific heat and ratio ofspecific heats vary as a function of temperature for differentequivalence ratios and EGR amounts.

Note that temperature dependence of the specific heatratio is quite significant. Based on the heat release equa-tion shown in (9), specific heat ratio acts a temperature-dependent nonlinear gain on the pressure transducer. This canparticularly be pronounced for advanced combustion modes


200 400 600 800 1000 1200 1400 1600 1800 20001.24

1.26

1.28

1.3

1.32

1.34

1.36

1.38

1.4

1.42

Temperature (K)

(ratio

ofsp

ecifi

che

ats)

= 1 + R/cv

AirPhi=1, EGR=60% (Pre-combustion)Phi=1, EGR=60% (Post-combustion)Phi=0.2, EGR=60% (Pre-combustion)Phi=0.2, EGR=60% (Post-combustion)

Fig. 4. Ratio of specific heats computed at different temperatures and mixturecombination (air-fuel ratio and EGR percentages). *: an example constant γvalue used during the entire combustion.

since the mixture temperatures at the time of fuel injectionpulses following main will be significantly higher than theinitial one because of prior heat release caused by the maincombustion.

To demonstrate the heat release computation with or withoutproperty tracking, test results from an advanced diesel combus-tion mode at 1800 r/min, 6-bar brake mean effective pressure(BMEP) are included where the fuel injection profile has twopulses: main and post.

Note that the heat release in Fig. 5 is relatively higherat the time of post when the variable mixture properties areproperly taken into account. This can be explained with Fig. 6.After main injection, the gas temperature increases, leadingto a drop in γ. Since a mixture with less γ requires morecombustion energy to generate the same pressure change, postheat release is higher than what would otherwise be computedwith a constant γ.

III. COMBUSTION PHASING CONTROL WITH

MULTIPLE-PULSE FUEL INJECTIONS

In this section, we present a novel combustion phasingcontrol methodology applicable to combustions with multiple-pulse fuel injections. The main objective of combustion phas-ing control is to align the entire heat release profile overthe crank-angle as desired. Due to the limited number ofcontrol variables to be adjusted, this objective is translated intomaintaining a key parameter from the heat release profile at adesired value. The most common way to control combustionphasing is to determine the 50% mass fraction burned point,which is practically the crank-angle denoted by CA50 at whichhalf of the fuel is burned (fuel energy is released), and to adjustthe start of main injection timing to maintain CA50 at a targetlocation.

In the case of more than one injection pulse, CA50 is stillgoing to give the 50% location of the overall heat releasedduring a combustion event, which will be somewhat affected

-80 -60 -40 -20 0 20 40 60 80 100-200

0

200

400

600

800

1000

1200

1400

1600

1800

Crank Angle (deg)

Hea

trel

ease

d(J

o ule

s)

Heat Release Profile

Hea release with variableHeat release with fixed =1.35

Fig. 5. Heat release computation from the same pressure data, with or withoutproperty tracking. Two injection-pulses (main and post) per combustion eventare used at this operating point as shown.

-80 -60 -40 -20 0 20 40 60 80 1001.3

1.31

1.32

1.33

1.34

1.35

1.36

1.37

1.38

1.39Change in value during combution cycle

CA (DATCD)

Fig. 6. Ratio of specific heats during combustion. The effect of main injectionon the properties during the post combustion is clearly notable by a drop inthe ratio of specific heats slightly after TDC because of main heat release.

by all the individual injection pulses. Therefore, by itself itmay not be sufficient to reveal which injection pulse needsto be corrected. This is more pronounced for combustionevents with main and post injections since both pulses willlead to significant heat releases at different sections of theoverall combustion. With this motivation, Fig. 7 illustratesthe basic architecture of the proposed phasing control sys-tem, which aims at extending the phasing control to havingindividual injection timing adjustment capability. As shownin Fig. 7, cylinder pressure from a particular combustionevent is processed to extract phasing metrics related to thedominant individual pulses. Phasing control computes theerrors between each metric and its operating-point-dependenttarget and adjusts the timing of the corresponding injectionpulse in closed loop to drive the error to zero.


Fig. 7. Base scheme of combustion phasing (timing) control.

Fig. 8. Heat release integral for three combustion events: same main injectiontiming, post injection timing varied.

In general, the set-point map for the combustion phasingand the nominal injection timings are referenced using enginespeed and load information (fuel command). Additional cor-rection can eventually be calibrated to compensate for intaketemperature, ambient pressure, fuel type, and the like. Baselinecombustion phasing set-points represent desired combustioncharacteristics with respect to plurality of criteria optimizingemissions, combustion noise, and fuel economy.

As for the detection of individual phasing, the first choicecould be to define another mass fraction burned percentage,as an example percentage, CA80 for example, to get the samenumber of feedback signals as the number of injection pulses.To demonstrate the results, we present several computed heatrelease traces from combustion with pilot, main and postinjection in Fig. 8. Specifically, three different combustionevents are plotted at the same operating point where the pilotand main injection timings are the same for all cases whilethe post injection timing is varied.

Note that all percentage mass fraction burned points arecoupled because of normalization at the end of combustion.In other words, for the case in Fig. 8, both CA50 and CA80 willbe affected by main and post injection timings and one cannot

Fig. 9. Heat release rate for three combustion events: same main injectiontiming, post injection timing varied.

easily associate the variations in the measured percentagepoints with the variations in the individual fuel pulse injectiontimings. Furthermore, if we choose to make bulk adjustmentsfor the entire injection profile for simplicity (like shifting mainand post-timing together based on CA50 only), this will resultin a closed loop control correction for the main injectiontiming, although the heat release variations of this exampleare caused by different post injection timings only.

Fig. 9 shows the heat release rates (before integration) forthe same data. It can be noted that the main heat release ratepeak location is not affected by the post injection, which startslater in the combustion cycle. Furthermore, the timing changesin the post injection pulse can easily be sensed from the peaklocations corresponding to the post combustion. This impliesthat the peak locations of heat release rates provide moreindependent information on the combustion phasing. For thecurrent application, a global peak algorithm was sufficient toextract main heat release peak location information. On theother hand, we employed a local peak detection algorithm forthe operating points with post injection. This also applies topilot injection, although active CL control of pilot injectiontiming was not one of the combustion control objectives forthis application. The combustion needs and associated controltasks related to pilot injection differ from main and postinjection, which are beyond the scope of this paper. It shouldbe noted that pilot injection (even multiple pilots) is verycommon in production applications. For the current detectionmethod, the primary consideration is how well the combustionevent of a given pulse is observable from the heat releaserate signal. Since pilot injection quantities are small, theircontribution to the heat release profile is also small, whichoften makes it relatively harder to sense it reliably from thecomputed heat release rate. Nevertheless, the existence of pilotinjection does not affect the detection of other major pulsesadversely. If the injection strategy has more than two pulses,yet the pulses are torque-forming and distinct, then the samestrategy can still be applied.


Fig. 10. Engine r/min and total injected fuel quantity (per stroke) from atransient run. In the fuel quantity trace, the section where two injection pulses(main and post) are employed is highlighted.

For peak location detection, the computed heat release ratewas first filtered with a moving average filter with variablecoefficients to suppress possible noise in the data. Since thecylinder pressure trace is available at the end of combustion,the moving window was placed around the current locationby including five sample points (two before the current point,two after the current point, and the current point itself) intotal. For the local peak location detection corresponding to thepost injection, a precalibrated crank range, which is scheduledwith respect to the detected global peak location correspondingto the main heat release rate, was searched over. A smallersize window was moved inside the search range to check thecriteria of local maximum within this smaller window. If apossible point for local maximum was identified, its magnitudewas saved for comparisons with the results from other sectionsinside the post combustion peak search window. A large setof cylinder pressure traces gathered at transient runs were firststreamed offline to calibrate and validate the peak detectionlogic and then validated further during closed loop controlexperiments.

Figs. 10 and 11 show the discussed combustion-phasing-related metrics from a transient run. Note that the heat releasefrom a single combustion describes what is happening during acombustion cycle over the crank-angle like in Fig. 9. The RCP-based pressure processing system generates the percentagemass-fraction-burned and peak heat release rate locations foreach combustion event. Figs. 10 and 11 show the computedmetrics as a function of time where new values come at everycombustion event. It is seen that there is a sudden change inCA50 when post injection starts. Specifically, CA50 moves inthe retard direction because of the late heat addition of postcombustion. On the other hand, the peak location of mainheat release rate is not influenced by post and can be used togauge variations in the main injection timing as a feedbacksignal. Furthermore, the peak location of post combustionprovides additional information on the post injection timingas a feedback measure.

190 200 210 220 230 240 250 260 270 280

-40

-20

0

20

Specific burn profile locations and main timing

Time (sec)

DBTD

C

190 200 210 220 230 240 250 260 270 280

-40

-20

0

20

Time (sec)

DB

TDC

CA10CA20CA50SOI: main timing

Main HRR max locationPost HRR max locationSOI: main timing

CA50 with post

Fig. 11. Computed % mass fraction burned (C A10, C A20, and C A50), peakheat release rate locations, and main injection timing during the transient runof Fig. 10. All the angles are in degrees before TDC.

0 10 20 30 40 50 60 70 80 90-15

-10

-5

0

Time (sec)

Mai

nP

eak

HR

Rlo

c(D

BT D

C) Phasing control with post

0 10 20 30 40 50 60 70 80 90-35

-30

-25

-20

Time (sec)

Post

Peak

HR

Rlo

c(D

BTD

C)

Fig. 12. CL control demonstration with main and post injection timings(main and post peak heat release rate locations).

An example CL combustion phasing control performance isshown in Figs. 12 and 13. For this case, the objective was tomake the main combustion track a desired profile while thepost combustion location is kept the same (see Fig. 12). Inorder to achieve this objective, the control looks at main andpost peak heat release rate locations and individually adjustsmain and post injection timings around their baseline values.For feedback corrections, a standard proportional, integral, andderivative (PID) in combination with a low-pass first-orderfilter was tuned in terms of steady-state accuracy, transientresponse, and noise rejection. A separate controller of thesame structure was applied to each individual cylinder andinjection pulse. The multiple lines in the figures correspondto individual cylinder measurements and control adjustmentsthrough cylinder-specific controllers (from an eight-cylinder


0 10 20 30 40 50 60 70 80 90-4

-2

0

2

4

Time (sec)

Mai

ntim

ing

offs

ets

(deg

)Main-post phasing control mechanization

0 10 20 30 40 50 60 70 80 90-5

0

5

Time (sec)

Pos

ttim

ing

offs

ets

(deg

)

Fig. 13. CL control demonstration with main and post injection timings(feedback corrections to main and post injection timings).

application). Note that applying the same timings for eachinjector does not guarantee the balanced combustion per-formance (see the larger cylinder-to-cylinder variations priorto control). Instead, the feedback control provides cylinderbalancing, it trims each individual cylinder main and post-timings so that the resulting combustion performances are asdesired and the same for all cylinders (see Fig. 13 wherethe final timings for each cylinder end up being different tomaintain the same combustion performance for each cylinder).

IV. FUEL BALANCING WITH MULTIPLE-PULSE

FUEL INJECTIONS

In this section, we address the problem of load imbalanceamong cylinders which can arise when the injected fuelquantities for each cylinder differ from each other. Fuelingvariations among cylinders is a common issue that can result intorque and speed fluctuations perceptible to driver. In practice,crank-shaft speed can be analyzed to detect and trim individualfuel amounts for a balanced torque delivery. However, thisapproach has limited accuracy due to limited bandwidth ofthe crank-shaft speed measurement and driveshaft torsionalvibrations. In the case of cylinder pressure sensing, IMEPcan be used as a load indicator to gauge individual cylindertorque contributions and adjust the fuel injection amounts inCL accordingly to get the same torque from each cylinder.For combustions with only single-pulse fuel injection, thebalancing algorithm can generate a delta fuel quantity for eachindividual cylinder and trim the corresponding cylinder’s mainfuel pulse-width accordingly to generate the same IMEP fromeach cylinder. In the case of multiple-pulse fuel injections,a new question arises of whether the fuel amount correctionshould be applied to main injection only or split in some wayamong the individual pulses. To open up this question, weapply a small signal variation analysis to the load metric,IMEP, by considering a particular combustion with main and

Fig. 14. Base scheme of fuel-balancing control.

post injections as follows:IMEP =

∫P(dV/V )

= f (Qmain,Qpost)

IMEP ≈ ∂ f

∂ Qmain�Qmain + ∂ f

∂ Qpost�Qpost

+ f (Q̄main, Q̄post)

IMEP ≈(

∂ f

∂ Qmaing + ∂ f

∂ Qpost(1 − g)

)�Q

+ f (Q̄main, Q̄post). (11)

In (11), �Q represents the total fuel amount variationgenerated by the balancing controller to make a particularcylinder’s IMEP the same for all cylinders. Depending on theload contribution of main and post injection pulses, there isan additional degree of freedom in splitting this delta fuelbetween the main and post injection pulses. With split factor“g” being one, the control action reduces to balancing IMEPwith only main fuel pulse-width adjustment as in the currentalgorithm. Alternatively, one can also use the original splitratio at the baseline calibration, e.g., g = Q̄main/(Q̄main +Q̄post). Although the load balancing can be achieved withdifferent split ratios, we exploit this additional degree offreedom to achieve the balance by finding out the requiredindividual fuel correction amounts for individual pulses.

Fig. 14 shows the proposed control system, which extendsthe fuel and load balancing controller with a mechanismof independently controlling the individual torque-formingpulses. In the sequel, we use the “load balancing” and “fuelbalancing” terms interchangeably since the control is alwaysacting on the fuel amounts in the current setting.

As for detection of individual fuel pulse amount variations,the main idea is to sample a fuel-mass-burned or load-relatedmetric sequentially along with injection pulses. Toward thisgoal, we first plot a representative cumulative PdV integralover crank-angle using a pressure trace from an operatingpoint at 2000 r/min and 7-bar BMEP in Fig. 15. For thisparticular combustion, injection profile includes two pulses;the main and the post, where the start of injections are 4°before TDC and 14° after TDC, and the quantities are 36.5and 16 mm3, respectively. Note that, the IMEP integral startsat −180° and its final value at +180° is the IMEP for this


-150 -100 -50 0 50 100 150

-10

-5

0

5

10

Crank Angle (deg)

-180

PdV

/Vd

( bar

)

Fig. 15. Cumulative IMEP integral is plotted. The integration starts at−180° and the value of the signal at 180° is the IMEP for the cycle. Forthis combustion, the fuel injection profile has two fuel pulses: main and postinjections.

Fig. 16. Metric for fuel balancing is demonstrated on a fuel burned plot(scaled heat release integral plot). For this combustion, the fuel injectionprofile includes two fuel pulses: main and post injection.

cycle. From Fig. 15, it is apparent that IMEP integral firstgoes negative because of work done by the piston duringcompression and reaches a positive value in the power stroke.Although main and post injections contribute to the final value,the instantaneous load is masked by the motoring pressureeffects, which would only cancel out once the PdV integrationis completed.

Therefore, the heat release has been studied as an alternativefuel burned mass metric. Fig. 16 shows the burned fuel masstrace from the same pressure data, which is approximatedby dividing the heat release with the lower heating value ofthe fuel. Two samples from the burned fuel mass trace inFig. 16 are highlighted. One sample is taken just before postcombustion starts and the second sample is taken at the endof entire combustion. Note that the sampled value prior to

Fig. 17. CL individual cylinder fuel-balancing control demonstration formain and post injection quantities (total fuel burned and fuel burned prior topost injection).

Fig. 18. CL individual cylinder fuel-balancing control demonstration(feedback corrections to main and post injection quantities).

post is not going to have any effect from the post injectionwhereas the final sample includes contribution from eachpulse.

An example of fuel-balancing control response is shownin Figs. 17 and 18. The results are shown before and afterthe activation of fuel-balancing control. In Fig. 17, both thetotal fuel burned and the fuel burned prior to post injectionare plotted from each cylinder. Baseline variations among thecylinders are very apparent prior to the activation of fuel-balancing controller. Fig. 18 shows the feedback correctionsto main and post injections for each cylinder. Once thecontroller is activated, main and post injection pulse-widthsof each cylinder are corrected by the feedback controller sothat the fuel burned samples prior to post combustion and atthe end of combustion are, respectively, balanced among thecylinders (see Fig. 17). For balancing, the moving averagevalue of the burned fuel measurement from a number of


Fig. 19. Engine r/min and total injected fuel quantity (per stroke) froma transient run as well as IMEP values from individual cylinders: Thesection where two injection pulses (main and post) are employed and thecorresponding IMEP imbalances are highlighted.

combustion cycles is computed for each individual cylinder.A feedback error is generated between this average and aparticular cylinder’s measurement. For feedback corrections,a dynamic regulation controller (with integral action) realizedby PID in combination with a low-pass first-order filter wasused, as in the case with phasing controls. Since the baselinefuel is requested by the driver, the fuel-balancing controllerfurther aims at maintaining the total fuel correction at zeroso as not to interfere with the driver torque request. For thiseight-cylinder application, this is achieved by using sevenindependent feedback loops per pulse per cylinder (“mainfuel quantity balancing control” block and “post fuel quantitybalancing control” block) and computing the eighth cylindervalue to make the total feedback correction zero for each pulse(“main-post injection coordination”).

The performance of the fuel-balancing control during anFTP transient segment is shown in Figs. 19–21. In Fig. 19,engine r/min and fuel amounts are shown along with thecomputed IMEP values for each cylinder without fuel balanc-ing. The sections at which the post injection is employed are

Fig. 20. CL individual cylinder fuel-balancing control demonstration duringtransients (total fuel burned and fuel burned prior to post injection).

Fig. 21. CL individual cylinder fuel-balancing control demonstration duringtransient (feedback corrections to main and post injection quantities).

highlighted (in the fuel plot, red trace: post fuel amount, greentrace: main fuel amount, blue trace: total fuel amount). Notethat IMEP imbalance deteriorates more when post injection isturned on.

Figs. 20 and 21 show the closed loop fuel-balancing controlresponse during the same segment of the FTP. Fuel-balancingcontroller balances the total fuel burned (samples at the endof combustion) and the samples prior to post (denoted bypartial load in Fig. 20) among cylinders by perturbing the mainand post injection pulse-widths for each cylinder automatically(Fig. 21). Note that fuel-balancing controller for post injectionis triggered when the post injection is used. Fuel-balancingcontroller for main injection is always on except a shortinterval around 370 and 405 s where the engine was not beingfueled. For the current engine calibration, we were able to use


Fig. 22. Individual cylinder IMEP values with fuel-balancing control fromthe same transient run of Fig. 20.

330 340 350 360 370 380 390 400 410

-10

-5

0

5

Time (sec)

Mai

nP

eak

HR

Rlo

c(D

BTD

C) Peak HRR locations

330 340 350 360 370 380 390 400 410

-35

-30

-25

Time (sec)

Post

Peak

HR

Rl o

c(D

BTD

C)

Main HRR peak locations

Post HRR peak locations

Fig. 23. CL individual cylinder combustion phasing control demonstrationduring transients (peak heat release rate locations corresponding to main andpost injections).

the same sampling locations over the entire transient run (15°after TDC and 90° after TDC for this run).

Fig. 22 shows the IMEP for each cylinder during fuel-balancing control. Note that the control did not use IMEP,but instead used the fuel burned samples from the heat releaseintegral. Nevertheless, fuel-balancing control still achieves thebalancing of the IMEP among cylinders as well.

The combustion phasing control was also active during thistransient along with fuel-balancing controller. The results areshown in Figs. 23 and 24. As detailed before, individualinjection timings are adjusted on the basis of the detectedheat release peak locations. The peak locations are plotted inFig. 23 as computed by CPPU and the corresponding timingfeedback corrections are shown in Fig. 24. As in the fuel-balancing controller, the timing controls associated with post

330 340 350 360 370 380 390 400 410

-0.5

0

0.5

Time (sec)

Pos

ttim

ing

offs

ets

(deg

)

330 340 350 360 370 380 390 400 410

-2

0

2

Time (sec)

Mai

ntim

ing

offs

ets

(deg

)

Individual cylinder timing corrections

Main injection timing corrections

Post injection timing corrections

Fig. 24. CL individual cylinder combustion phasing control demonstrationduring transients (feedback corrections to main and post injection timings).

injections are only triggered when the post injection is used,whereas the timing control associated with main injection isalways on except for the region where engine was not beingfueled at all.

V. OVERALL CYLINDER BALANCING PERFORMANCE

WITH HEAT RELEASE COMPARISONS

After the multiple-pulse fuel injection phasing and fuel-balancing controllers are designed and validated during steady-state and FTP transients, we look at their performance interms of overall heat release behavior. Specifically, individualcylinder heat release profiles are computed from a combustionevent with main and post injection at 1800 r/min, 6-bar BMEPfor three different cases: open loop (OL) and CL with differentfeedback measures.

Fig. 25 shows individual cylinder pressures and the corre-sponding heat release profiles along with the injection currentprofile from a sample cycle at this operating point withoutany CL corrections to the injection profile. In other words, theinjection timings and the pulse-widths are applied as calibratedfor each cylinder. The noticeable variations in the pressuretraces and the heat release profiles in Fig. 25 reflect thecylinder-to-cylinder baseline variations. For plotting purposes,the heat release profiles are normalized with their maximumvalues and scaled. The variability of individual cylinder CA50saround 20° after TDC can be seen in Fig. 25.

For comparison, Fig. 26 shows the same signals from thesame run with CL combustion phasing and load balancingcontrol using CA50 and IMEP as our baseline controllers.Particularly, individual cylinder CA50s are used to adjust theinjection timings and IMEP is used to trim fuel pulse-widths tobalance IMEP among cylinders. Since the control uses overallcombustion-related metrics, the corrections are applied as awhole: phasing control moves the entire injection profile andthe load balancing control corrects the main injection pulse-widths. As can be seen from Fig. 26, CA50s are balanced and


-20 0 20 40 60 80-2

0

2

4

6

8

10

12x 106

Crank angle (CA)

Pre

ssur

e(k

Pa)

and

scal

edm

ass

fract

ion

burn

ed1800 RPM, 6 bar

Fig. 25. Individual cylinder pressure traces and the corresponding heat releaseprofiles. The same main and post injection timings are applied. Individualcylinder variations are apparent through heat release profiles.

-20 0 20 40 60 80-2

0

2

4

6

8

10

12x 106

Crank angle (CA)

Pres

sure

(kP

a)a n

ds c

aled

mas

sf ra

ctio

nbu

rne d

1800 RPM, 6 bar

CA50 location

Fig. 26. Individual cylinder pressure traces and the corresponding heat releaseprofiles. As an intermediate case, individual cylinder CA50s are balanced byadjusting main and post injection timings together. IMEPs are also balancedby CL adjustments to main fuel quantities.

IMEPs are also balanced, although the data does not explicitlyshow the IMEP values. It is noticeable that balancing IMEPand CA50 without the account of the multiple injection-pulsenature of the combustion can lead to increased variations inthe cylinder pressures and heat release profiles in comparisonof Fig. 26 with Fig. 25.

Finally, Fig. 27 shows the results with the current combus-tion phasing and fuel-balancing controls. The phasing controladjusts main and post injection timings on the basis of peaklocations of the heat release rates, and the fuel-balancingcontrols trims both main and post injection pulse-widths on thebasis of the fuel burned metrics. As it can be seen from Fig. 27,the entire heat release is very well controlled and balanced

-20 0 20 40 60 80-2

0

2

4

6

8

10

12x 106

Crank angle (CA)

Pre

ssur

e(k

Pa)

and

scal

edm

ass

fract

ion

burn

ed

1800 RPM, 6 bar

Fig. 27. Individual cylinder pressure traces and the corresponding heat releaseprofiles. For this case, individual cylinder peak heat release rates are balancedby adjusting both main and post injection timings with the proposed method.Nominal main and post injection pulses are also included in the figure withoutshowing CL adjustments as a reference.

among cylinders when the multiple-pulse nature of the fuelinjection profile is taken into account in control system. It isalso clear that the control assures the regulation and balancingof CA50 and IMEP despite the fact that those metrics are notdirectly used by the control system.

The performance of proposed CL combustion control fordispersion reduction has also been tested at selected keyoperating points. Dispersion refers to the possible variations inthe emissions or other performance variables from calibratedvalues in real-world operation. Tighter CL management ofcombustion inherently reduces the level of dispersion. Inthis paper, a possible variability in the field is simulatedby perturbing the injection profile and the cylinder EGRamounts. Particularly, after running as calibrated, variations tothe injection profile and EGR valve are applied. For variationsin the combustion retard direction, both main timing is retardedby 2° and the separation time between main and post injectionis increased by 2° (retards post-timing). Similarly, for thecombustion advance direction, both main timing is advancedby 2° and the separation time between main and post injectionis reduced by 2° (advanced post-timing). In addition, EGRamount changes within 2% are applied for each case. Notethat, increase in EGR and retardation of injection timing havesimilar effects on the heat release. They both cause the heatrelease to be delayed later in the cycle. Similarly, decrease inEGR and advancement of injection timing have similar effectson the release of heat. They both cause the combustion to startand progress earlier in the cycle. Therefore, pairing the pertur-bation directions as such allows compounding their effects onthe combustion. In addition, as all the other calibration valuesare scheduled on the basis of fuel command and speed, anytorque changing perturbation resulting in a modified baselinefuel amount to keep the same BMEP will also cause shifts inother calibrations.


TABLE I

SNAPSHOT OF EMISSIONS AND COMBUSTION VARIABLES AT 1800 r/min,

6 BAR FOR VARIOUS SCENARIOS

1800r/min, 6

barBaseline

OLWith

RetardedTimings

andHigherEGR

CLDuring

RetardedTimings

andHigherEGR

OL WithAdvancedTimingsand Less

EGR

CLDuring

AdvancedTimingsand Less

EGR

Main peakHRR

[after TDC(ATDC)]

8.5 13 8.5 5.5 8.5

Post peakHRR

(ATDC)26 33 26 20.5 26

CA50(ATDC)

25 31 26.4 21.4 26

HC (ppm) 73.6 198 70.4 28.9 70.9

NOx(ppm)

26.9 15 25.7 37.2 26.6

Smoke 1.56 1.22 1.66 2.29 1.61

CN (dBa) 89.8 86.8 89.4 90.2 89.4

As shown in Table I, retarded timings and increased EGRcause more hydrocarbon (HC) emissions and higher combus-tion variability. Similarly, for the case with advanced maintiming and shorter separation time between the pulses, bothNOx and smoke emissions increase. A slight rise in combus-tion noise (CN) is also noted. The reason for these changescan be explained as follows: as the separation time betweenmain and post injection reduces, the two previously distinctcombustion events start to merge imitating the case with asingle injection pulse with large pulse-width, leading to astronger main combustion (higher NOx and noise) and lesssoot burning time (more smoke). On the other hand, the closedloop corrections to the injection profile and EGR valve providevery similar combustion behavior and emission numbers to thebaseline.

Overall, the CL combustion controls along with the pre-sented phasing and fuel-balancing control features providereal-time combustion monitoring and correction capability,which leads to tighter combustion performance despite varia-tions. It has been observed that small changes in the injectionprofile and air/EGR amounts can cause high jumps in respec-tive emissions for high-EGR combustion modes. The reductionof emission dispersion could benefit significantly from closedloop combustion controls.

VI. CONCLUSION

A multiple injection combustion phasing and fuel-balancingcontrol system based on in-cylinder pressure feedback wasdeveloped. The control system used information on the heatrelease rates and fuel burned profiles to maintain a desiredcombustion through online adjustments to the injection profilefor individual cylinders. The control system was mechanizedin an engine dynamometer with a rapid prototyping con-troller and was demonstrated to successfully regulate and

balance the combustion phasing and load for each cylinderduring steady-state and transient runs. The effectiveness of thepresented features for dispersion management as a part of theoverall combustion control system was also studied at severaloperating points. The studies showed that different types ofvariations in the calibration inputs can be detected through thereal-time heat release analysis and compensated accordinglyby the proposed methodology.

ACKNOWLEDGMENT

The authors would like to thank C.-F. Chang, F. Matekunas,and D. T. French for their discussions and support for thispaper.

REFERENCES

[1] R. J. Hosey and J. D. Powell, “Closed-loop knock adaptive sparktiming control based on cylinder pressure,” J. Dyn. Syst., Meas., Control,vol. 101, no. 1, pp. 64–70, Mar. 1979.

[2] F. A. Matekunas, “Engine combustion control with ignition tim-ing by pressure management ratio,” U.S. Patent 4 622 939, Nov. 18,1986.

[3] Y. Kowomura, M. Shinshi, H. Sato, N. Takashi, and M. Iriyama, “MBTcontrol through individual pressure detection,” in Proc. SAE Congr., Nov.1988, no. 881779.

[4] J. Powell, “Engine control using cylinder pressure: Past, present, andfuture,” ASME J. Dyn. Syst., Meas. Control, vol. 115, no. 2, pp. 343–350, Jun. 1993.

[5] S. Leonhardt, N. Muller, and R. Isermann, “Methods for engine super-vision and control based on cylinder pressure information,” IEEE/ASMETrans. Mechatron., vol. 4, no. 3, pp. 235–245, Sep. 1999.

[6] R. Muller, M. Hart, G. Krotz, M. Eickhoff, A. Truscott, A. Nobel,C. Cavalloni, and M. Gnielka, “Combustion pressure based enginemanagement system,” in Proc. SAE Congr., 2000, no. 2000-01-0928.

[7] M. C. Sellnau, F. A. Matekunas, P. A. Battiston, C. F. Chang, and D. R.Lancaster, “Cylinder-pressure-based engine control using pressure-ratio-management and low-cost non-intrusive cylinder pressure sensors,” inProc. SAE Congr., 2000, no. 2000-01-0932.

[8] I. Haskara, G. G. Zhu, C. F. Daniels, and J. Winkelman, “Oncombustion invariants for MBT timing estimation and control,” inProc. Int. Combust. Eng. Division Fall Tech. Conf., Oct. 2004, pp.243–250.

[9] J.-O. Olsson, P. Tunestal, and B. Johansson, “Closed-loop control of anHCCI engine,” in Proc. SAE Congr., 2001, no. 2001-01-1031.

[10] A. Gangopadhyay, F. A. Matekunas, P. A. Battiston, P. G. Szymkowicz,J. Pinson, and G. Landsmann, “Control of diesel HCCI modes usingpressure-based timing metrics,” in Proc. FISITA World Autom. Congr.,Oct. 2006, pp. 1–9.

[11] M. Hasegawa, Y. Shimasaki, S. Yamaguchi, M. Kobayashi, H. Sakamoto,N. Kitayama, and T. Kanda, “Study on ignition timing control for dieselengines using in-cylinder pressure sensor,” in Proc. SAE Congr., 2006,no. 2006-01-0180.

[12] M. Beasley, R. Cornwell, P. Fussey, R. King, A. Noble, T. Salamon,A. Truscott, and G. Landsmann, “Reducing diesel emissions dispersionby coordinated combustion feedback control,” in Proc. SAE Congr.,2006, no. 2006-01-0186.

[13] H. Husted, D. Kruger, G. Fattic, G. Rippley, and E. Kelly, “Cylinderpressure-based control of pre-mixed diesel combustion,” in Proc. SAECongr., 2007, no. 2007-01-0773.

[14] I. Haskara, A. Gangopadhyay, P. A. Battiston, F. A. Matekunas,and P. G. Szymkowicz, “Simultaneous EGR correction and individualcylinder combustion phase balancing,” U.S. Patent 7 231 906, Jun. 19,2007.

[15] I. Haskara, A. Gangopadhyay, P. A. Battiston, F. A. Matekunas, and P.G. Szymkowicz, “Internal combustion engine exhaust gas recirculationcontrol,” U.S. Patent 7 231 905, Jun. 19, 2007.

[16] R. Graglia, A. Catanese, F. Parisi, and S. Barbero, “The new generalmotors diesel engine management system,” in Proc. MTZ Worldwide,vol. 72. Feb. 2011, pp. 40–45.

[17] J. B. Heywood, Internal Combustion Engine Fundamentals. New York:McGraw-Hill, 1988.


Ibrahim Haskara received the B.S. degrees inelectrical and electronics engineering and in physicsin a double major from Bogazici University, Istanbul,Turkey, in 1995, and the M.S. and Ph.D. degreesin electrical engineering from Ohio State University,Columbus, in 1996 and 1999, respectively.

He has been with General Motors, Warren, MI,since 2005, where he is currently a Staff Researcher,specializing in propulsion control systems withPropulsion Systems Research Lab, GM GlobalResearch and Development, Warren. From 1999 to

2005, he was a Control Research Scientist with Visteon Corporation, VanBuren TWP, MI. He has published more than 40 peer-reviewed journal andconference papers and holds 25 U.S. patents for various automotive appli-cations. His current research interests include control, modeling, estimation,optimization and diagnostics of automotive systems, particularly in the areaof advanced propulsion systems and engine control management, as well ascontrol of nonlinear systems and control theory in general.

Yue-Yun Wang received the Ph.D. degree in electri-cal engineering from Shanghai Jiao Tong University,Shanghai, China, in 1987.

He is currently a Staff Researcher leadingadvanced CIDI engine controls development inPropulsion Systems Research Laboratory, GeneralMotors Research and Development, Warren, MI.From 1995 to 2005, he was a Technical Advisorwith Cummins Engine Company, Columbus, IN,specializing in next generation power train controland diagnostics. Before joining the industry, he

conducted research and taught at several academic institutions, where hewas an Assistant Professor with Shanghai Jiao Tong University, an Alexandervon Humboldt Research Fellow with Duisburg University, Essen, Germany,a Visiting Scholar with Syracuse University, Syracuse, NY, and a SeniorResearch Associate with Ohio State University, Columbus. He has publishedover 100 publications and holds 50 U.S. patents and numerous internationalpatents for automotive applications. His current research interests includemodeling and simulation, model-based multivariable control, robust control,adaptive control, and detection and estimation with application to automotivepropulsion systems.

Documents

Cylinder Pressure-Based Combustion Controls for Advanced Diesel Combustion With Multiple-Pulse Fuel Injection