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Doctoral Thesis

Load control of SI engines using secondary valves

Author(s): Vogel, Olivier Denis

Publication Date: 2000

Permanent Link: https://doi.org/10.3929/ethz-a-003913266

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Diss.ETHNr. 13633

Load Control of SI Engines

Using Secondary Valves

A dissertation submitted to the

SWISS FEDERAL INSTITUTE OF TECHNOLOGY

ZURICH

for the degree of

Doctor of Technical Sciences

presented by

OLIVIER DENIS VOGEL

Dipl. Masch.-Ing. ETH

born 3.1.1969

citizen of Bonfol, JU

accepted on the recommendation of

Prof. Dr. L.Guzzella, examiner

Prof. Dr. D.Favrat, co-examiner

2000

Seite Leer /

Blank leaf

"A mind once stretched

by a new idea,

never regains its

original dimensions."

Anatole France

V

Acknowledgments

I wish to gratefully acknowledge the guidance and support of

Prof. Dr. Lino Guzzella, my advisor, during this research.

Furthermore I would like to thank Prof. Dr. D. Favrat, Director of the

Laboratoryfor Industrial Energy Systems, EPFL, Lausanne, for accepting to

be my co-examiner.

A special thanks goes to Dr. Kimon Roussopoulos who always encouragedme and without whose help this work would hardly have been accomplished.

I am also indebted to Dr. Alois Amstutz and to the entire staff of the

Laboratory of Engine Systems, especially to Dr. Chris Onder.

The list of people who helped me in some way in the course of this research

would be too long to be printed here. But I would like to assure all of them of

my gratitude.

Finally, my thanks go to my parents whose support and encouragement

during all the years made this dissertation possible in the first place.

VI

Contents

Acknowledgments v

Abstract ix

Zusammenfassung xi

Notations xiii

1 Introduction 1

1.1 Basic Ideas 1

1.2 Approach 6

1.3 Objectives of the Thesis 8

1.4 Main Results and Own Contributions 8

2 Cylinder Charge Control 11

2.1 Throttle Plate and Fixed Camshaft Timings 11

2.2 Variable Valve Actuation 12

2.3 Overview of VVT Devices 13

3 Secondary Valves 19

3.1 Application of Secondary Valves 19

3.1.1 Part Load Optimization 19

3.1.2 Roughness Improvements 25

3.1.3 Idle Speed Reduction 27

3.1.4 Cylinder Cut Off 27

\

vii

3.2 Benefits and Drawbacks of Secondary Valves 28

3.2.1 Benefits 28

3.2.2 Drawbacks 29

4 Models of Engines with Secondary Valves 31

4.1 Engine Model Requirements 31

4.2 Selection of an Engine Model 34

4.3 Model of the Secondary Valve and the Engine 36

4.3.1 Main Equations 36

4.3.2 Secondary Valve Characteristics 41

4.3.3 Gas Components and their Properties 42

4.3.4 Burn Rate Estimation 43

4.3.5 Wall Heat Transfer 48

4.3.6 Spray Fuel Evaporation and other Considerations 49

4.4 Validation of the Simulation 49

4.4.1 Conventional Engine 49

4.4.2 Engine Equipped with Secondary Valves 54

5 Main Results of the Simulations 57

5.1 Analysis with Willan's Approach 58

5.2 Fuel Map 61

5.3 Residual Gas Fraction Map 64

5.4 Burn Delay Map 65

5.5 Secondary Valve Timing Map 66

6 Model Sensitivity Analysis 69

6.1 Variation of the Parasitic Volume 69

6.2 Variation of the Manifold Temperature 73

6.3 Variation of Secondary Valve Closing Times and Durations 75

vin

7 Experiments 79

7.1 Static Measurements - Engine with Secondary Valves 79

7.1.1 Measurement Setup 80

7.1.2 Secondary Valve Prototype 80

7.1.3 Roughness Improvements 87

7.1.4 Load Control 90

7.2 Static Measurements - Conventional Engine 94

7.2.1 Measurement Setup 95

7.2.2 Result of the Static Measurements 95

7.3 Dynamic Measurements - Conventional Engine 95

8 A case study: MVEG-95 Fuel Consumption 97

8.1 Quasi-Static Simulation for Fuel Consumption Estimation 98

8.1.1 Main Equations 101

8.1.2 Validation 103

8.2 Comparison of Experiments and Simulations 109

8.2.1 Driving Cycle Simulation - Results 109

9 Conclusions and Outlook 111

Appendix

A Parameters 115

A.l Main values - engine 115

A.2 Main values - vehicle model 116

References 117

Curriculum Vitae 123

IX

Abstract

This thesis analyzes a novel approach for improving the efficiency of SI

engines. Conventional engines use a throttle to control the engine load, caus¬

ing substantial pumping losses at part load. An alternative way of implement¬

ing load control is to use an early intake valve closing (EIVC) scheme, which

can reduce pumping losses significantly. To implement this, the conventional

camshaft system is usually replaced with active devices (electromagnetic,

hydraulic or other) which permit the variation of intake and exhaust valve tim¬

ings. This thesis, in contrast, investigates a new system with similar impact on

the engine cycle: by using additional secondary valves in the intake runner,

engine load can be controlled without modifying the conventional camshaft

system. These new valves can be made of lightweight plastic material which

will permit fast transition durations (of the order of milliseconds) and more¬

over require low actuation power. The applications of such valves are not lim¬

ited to load control, they also enable the reduction of internal residual gas

recirculation at part load conditions. With that it becomes possible to use cam¬

shafts with large valve overlap enabling increased peak power even at low

engine speed since the unacceptably high engine roughness of an aggressivecamshaft at part load is reduced significantly.

An important aspect of this work is to estimate the fuel consumption of a

vehicle during a driving cycle by using two models: the first captures the ther¬

modynamic engine cycle and computes, among other results, fuel consump¬

tion for stationary engine conditions. With these results, part load fuel

consumption maps are derived which form the base for the second simulation.

The second model computes the fuel consumption of a vehicle during a driv¬

ing cycle. Two different engine case-studies are analyzed: a vehicle with a

conventional engine, and a vehicle whose engine is equipped with secondaryvalves. The final results show that in the latter case a fuel consumption reduc¬

tion of 6.2% can be achieved during a New European Driving Cycle (MVEG-

95) test when the power to drive the secondary valve actuators is neglected.

X

Both models are validated with experiments on a test bench. These experi¬ments are performed on a Ford Zetec DOHC 4-cylinder 2.0-liter engine. Once

with a conventional throttle and once with secondary valves on each cylinder

port.

xi

Zusammenfassung

Die vorliegende Arbeit befasst sich mit dem Einsparungspotenzial des Kraft¬

stoffverbrauchs eines Viertakt Ottomotors beim Einsatz eines neuartigen Last¬

steuerungssystems. Bei konventionellen Motoren wird die Last mit einer

Drosselklappe gesteuert, welche im Teillastbereich erhebliche Pumpverlusteverursacht. Eine andere Möglichkeit der Laststeuerung besteht im Konzeptdes frühen Einlass schliessen (FES), welches den Vorteil aufweist, die Pump¬verluste beträchtlich zu reduzieren. Dazu wird im allgemeinen der konventio¬

nelle Nockenwellentrieb der Ein- und Auslassventile durch einen

elektromagnetischen oder hydraulischen Antrieb ersetzt, damit die Ventilsteu¬

erzeiten variabel eingestellt werden können. In dieser Arbeit wird ein neues

System vorgestellt, welches eine ähnlichen Effekt auf den thermodynami-schen Motorprozess ausübt: Zusätzliche Ventile im Einlasstrakt übernehmen

die Laststeuerung, ohne den konventionellen Nockenwellentrieb der Ein- und

Auslassventile zu ersetzen. Die aus Kunststoff hergestellten zusätzlichen Ven¬

tile müssen im Bereich von Millisekunden schliessen oder öffnen, um effizi¬

entes FES zu ermöglichen. Der Einsatz solcher zusätzlicher Ventile

beschränkt sich nicht nur auf die Laststeuerung, sondern ermöglicht auch eine

Beeinflussung der innermotorischen Abgasrezirkulation. Damit können

Nockenwellen mit grossen Ventilüberschneidungen verwendet werden, wel¬

che bei konventionellen Motoren im Teillastbereich zu nicht akzeptablenMotorvibrationen führen würden. Das zweite Ziel dieser Arbeit ist die Kraft¬

stoffverbrauchsabschätzung eines Fahrzeugs in einem Fahrzyklus. Dazu wur¬

den zwei Simulationen verwendet. Die Simulation des thermodynamischen

Motorprozess liefert unter anderem den Brennstoffverbrauch in einem statio¬

nären Betriebspunkt. Durch die Berechnung verschiedener Betriebspunktewurden Verbrauchkennfelder ermittelt, welche die Basis für die zweite Simu¬

lation bilden. Diese berechnet den Kraftstoffverbrauch eines Fahrzeuges wäh¬

rend eines Fahrzyklus, wobei zwei Fälle untersucht wurden: Fall a) für einen

konventionellem Motor, Fall b) für einen mit zusätzlichen Ventilen ausgerü¬

steten Motor. Die Resultate zeigen, dass im Neuen Europäischen Fahrzyklus

Xll

(MVEG-95) im Fall b) ein Kraftstoffeinsparpotenzial von bis zu 6.2% vor¬

liegt. Der Energiebedarf der Ventilaktuatoren wurde in der Berechnung nicht

berücksichtigt.

Beide Simulationen wurden am Prüfstand validiert. Die Untersuchungenwurden an einem Ford DOHC 4-Zylinder Motor durchgeführt, wobei die

Laststeuerung wahlweise mit einer Drosselklappe oder mit zusätzlichen Venti¬

len erfolgte.

xiii

Notations

1. Symbols

SI units (Système International d'Unités) if not specified

otherwise in the text.

a Vibe parameter

r\ Efficiency

A Area

B Woschni constant

c Specific heat

cd Air friction coefficient

c Specific heat at constant pressure

c, Rolling friction coefficient

cy Specific heat at constant volume

C Woschni constant

E Energy

f Csallner weighting parameter

gGravitational acceleration

Csallner weighting parameter

h Specific enthalpyCsallner weighting parameter

h Heat transfer coefficient

H Enthalpy

i gear ratio

XIV

m Mass

Vibe parameter

Woschni parameter

m Mass flow

mab Fuel mass burned

mf Fuel mass flow

n Crankshaft rotational speed

Polytropic exponent

nr Number of crank revolutions per power stroke

P Pressure

P Power

WP Work transfer of piston

Q Heat transfer

Qhv Fuel heating value

rw Wheel radius

Rf Fuel relevance

SP Piston speed

T Temperature

u Specific internal energy

V Velocity

V Volume

Vd Displaced cylinder volume

w Work

Work transfer

X Residual gas fraction

XB Burn fraction

XB Burn rate

e Crank angle

% Crank angle at end of burning

*i Crank angle at ignition

e, Crank angle at start of burning

0 Inertia

X Relative fuel-air ratio

Ç Residual gas fraction

p Density

9 Crank angle

CO Angular velocity

2. Subscripts

0 Section - intake runner entry

Reference value

1 Section - secondary valve

2 Section - intake valve

3 Section - exhaust valve

01 System in front of SV

12 System behind SV = parasitic volume

23 System - cylinder

P Parasitic volume

am Air-fuel mixture

at Total air in cylinder

aux Auxiliaries

ht Total burned gas in cylinder

B Burning

cClutch

Clearance

corr Corrected

XVI

cy or C Cylinder

e Engine

i ignition

meas From measurements

n Crankshaft rotational speed

P Pressure

PI Pumping loop

r Residual gas

ref Reference point

rel Relative

sim From simulation

SV Secondary valve equipped engine

t Transmission

thr Throttled engine

T Temperature

w or W Wall

X Residual gas fraction

X Relative fuel-air ratio

3. Notation

A Difference

Average or mean value

Flux or flow

d

dtDerivative in time

d

dQDerivative in crankshaft degrees

xvii

4. Abbreviations

bmep Break mean effective pressure

fmep Friction mean effective pressure

fuel mep Fuel mean effective pressure

imep Indicated mean effective pressure

mep Mean effective pressure

ATC After top dead center

BDC Bottom dead center

CA Crank angle

ECE Urban driving cycle

EUDC Extra-urban driving cycle

EVC Early valve closing

MBT Maximum break torque

MVEG-95 Motor Vehicle Emissions Group

New European Driving Cycle

QSS Quasi static simulation

STD Standard deviation

SV Secondary valve

TDC Top dead center

VIVC Variable intake valve closing

VVT Variable valve train

1

Chapter 1

Introduction

Modern engines and vehicles are the result of over a century of research and

development. The performance profile demand was constantly changing but

converged mainly towards a search for a powerful yet environmentally

friendly engine. Bringing both aspects together was and remains the work of

engineers, who have to predict and realize future customer demands whilst

satisfying forthcoming legal restrictions.

This work shows one approach how the fuel consumption of a vehicle duringa driving cycle can be estimated. Starting with a novel concept for engine load

control, engine cycle analysis is performed demonstrating all main and sec¬

ondary benefits of the concept. These results allow a detailed vehicle simula¬

tion which finally, together with a given driving cycle, provides the total fuel

consumption close to real conditions.

This set of the engine cycle model, on the one hand, and the vehicle simula¬

tion on the other hand offers a quick and convenient method of estimating fuel

benefits of almost every new component or concept independently if the

engine process or vehicle is affected.

1.1 Basic Ideas

Production engines are usually optimized for maximum power which by the

nature of the process is developed at high speed and torque. During normal

driving condition this power is almost never used. Despite peak engine effi-

2 Chapter 1 Introduction

at a Specific Speed

Figure 1-1: Willan's line of a typical SI engine at fixed engine speed.

ciencies of T| = 0.32...0.38 the average efficiency under normal driving condi¬

tions drops dramatically to r\ = 0.1...0.2, depending on installed engine powerand vehicle properties. Different aspects of this problem have been analyzedand one of the most promising solution is the reduction of installed power.

Taking into account that the customer is not ready to accept this solution other

ways of offering better part load efficiency without reducing peak power need

to be investigated.

Engine efficiency can be expressed graphically using the Willan's plot whichassumes an affine relation between break mean effective pressure and fuel

mean effective pressure at a fixed speed as shown in Figure 1-1

([1], Chapter 13). Both useful relative engine performance measures, obtained

by dividing the work per cycle, or released fuel energy respectively, by the

cylinder volume displacement.

P-nrbmep = —

(1.1)Vd'n

1.1 Basic Ideas 3

rrif QHV-nrfuelmep = (1.2)

Vd-n

Thus engine efficiency is

Tl = ^- (1.3)fuel mep

Iso-efficiency curves therefore starts from zero bmep and fuel mep and are

straight lines.

Great efforts have been made to increase peak engine efficiency, which

results in lifting the end of the Willan's line (region II) to higher bmep and any

solution were conceived and realized to achieve improved peak engine power- for example intake runner length optimization for ram effects on cylinder

charge, increased charge motion for faster burning or variable valve timing for

enhanced scavenging. The impact of these enhancements have often no or lit¬

tle influence in part load operating points; the regime the engine is mainlyused.

Lifting the lower end of the Willan's line, that is to increase bmep0^ is

another approach to heightening engine efficiency. The value bmep0 repre¬

sents mainly engine friction and pumping loop work of a motored engine

(region I) but includes energy used for auxiliaries as well. The reduction of

any of these losses will increase engine efficiency of part load and needs to be

considered. Friction losses can hardly be changed without modification of the

engine block and bearings or enhancing oil properties whereas pumping looplosses offers strong opportunities of rising bmep0. Inasmuch as up to 50% of

the total losses are generated by throttling the inlet charge before entering the

cylinder [2] [6]. Auxiliaries are not addressed in this work but should not be

neglected since 5 to 10% of fuel consumption can be saved by improving their

design and energy management [3].


Because typical engine operating points lie in the first half of the fuel mepaxis range, low load solutions to increase bmep0 should be the focus of further

investigations. The best known techniques for reducing pumping work are:

1. variable valve timing (VVT)2. charge heating3. lean burning4. downsizing and supercharging

This work focuses on variable valve timing and to some extend also on

charge heating - the importance of the other two points is acknowledged but

an overall analysis would go beyond the scope of this thesis.

In conventional automotive engines the inlet and exhaust valves are driven

mechanically with a direct and essentially fixed connection to the enginedrive-shaft. For this reason the timing of the intake and exhaust events is

invariant with engine speed and load: clearly it cannot be optimal at all condi¬

tions, and the timing therefore becomes a compromise. A great deal of

research has been and is being undertaken into means of implementing vari¬

able valve timing schemes, which enable the valve timings to be varied as a

function of engine speed and load. Simple camshaft phase-shifting schemes

are offered in several commercially available passenger cars.

Fully flexible electro-magnetic or electro-hydraulic VVT systems have been

presented by several groups .No commercial vehicle has however been

equipped with such systems, yet. These VVT systems have some fairly severe

functional requirements.

Secondary valves operate in a similar way but without being exposed to the

combustion chamber: the principle is that secondary valves act in series with

the conventional poppet valves of the engine Figure 1-2, which are actuated in

the conventional manner and with fixed timing. Secondary valves are not

required to withstand the temperatures and pressures of combustion, and

therefore can be made of light-weight plastic material. It is thereupon easier to

adjust the timing of the secondary valve than that of the main valve. The con-

1. see Chapter 2

1.1 Basic Ideas

Position of the

Secondary Valve

Parasitic Volume

m ~r~rr

rr-r 3=rt

1,11 1 II1,1

o

Figure 1-2: Secondary valve: placement

cept proposed in this thesis is to use secondary valves in conjunction with

main valve timings optimized for high torque at high engine speed which

implies large timing overlaps. Secondary valves would be used at low speed to

control internal exhaust gas recirculation (EGR) and at low loads also for cyl¬inder charge control, i.e. load control. At high engine speed they would be

deactivated. In conjunction with aggressive camshafts, offering a better scav¬

enging of the cylinder and thus a better engine charge, the peak power of the

engine can be increased by up to 15%.

Figure 1-1 shows at fuel mep = x the generated bmep when the engine is

operated with a) a throttle valve, b) secondary valves. The increased output at

low load is the primary motivation of implementing secondary valves on a SI

engine. Hence, fuel consumption benefits of secondary valves during a drivingcycle reflecting the common use of a car in urban environments remains the

final step of evaluating these new devices. Whether finally the engine is down¬

sized to attain the same power with less fuel consumption or left at the same

size such to produce more power output at same fuel consumption will not be

addressed here.


Fuel savings through pumping loss reduction and roughness improvementsare not the only possible benefits from throttleless control. The ability to con¬

trol the load from cycle to cycle and from cylinder to cylinder is in itself a

valuable approach in transient control and operating parameter optimization.Of course, these advantages would be all available from a secondary valve

system, too.

As mentioned above, several research teams have implemented camless vari¬

able valve timing systems that control the main engine valves. However the

design requirements for a secondary valve system are less severe in terms of

pressure or temperature resistiveness. Any actuation method suitable for

replacing the cam system should be suitable for operating a secondary valve

system - although the weight and power consumption of the system would be

additional to the conventional valve train rather than instead of it. One benefit

of a secondary valve scheme over a "camless engine" scheme is that, with

suitable "fail-safe" (i.e. fail open) design, the engine should still be able to

run, with reduced "home" functionality, even in the event of a secondary valve

failure.

1.2 Approach

Starting from the main idea of implementing early valve closing using fast

secondary valves, a simulation of the engine with this new process is carried

out. The engine model allows variations of all salient parameters of secondaryvalves such as closing speed, effective intake flow cross areas, etc. and enables

the simulation of a conventionally throttled engine as well. The analysis of the

results yields secondary valve timings, expected fuel savings and residual gas

reductions to name the most important results. The validation of the model

remains a key point and makes the design and implementation of a secondaryvalve test system necessary. A simulation of a vehicle during a driving cycle is

used to estimate the total fuel consumption. The validation of this second sim¬

ulation is done by emulating vehicle, driver and road on a dynamic test bench

with a production line engine, without any modification of the engine and its

control schemes. An overview of the chosen approach is shown in Figure 1-3.

1.2 Approach 7

SecondaryValves I

ThermodynamicModel of the

SI Engine

Engine PropertiesDimensions

Masses

# of CylindersValve TimingNew Devices

etc.

Gas PropertiesEnthalpiesHeating Values

Gas Constants

etc.

Physical Laws

Combustion Model

Burning Model

Backflow Model

etc.

Validation

Static Test bench

Measurements

Quasi-Static

Simulation

of a Vehicle

during Driving Cycle

Vehicle PropertiesDimensions

Mass

Drive Train

etc.

Driving Cycle PropertiesDistance

Given Gear

Given Speedetc.

Result

Fuel Consumption Map..

^Residual Gas Map

r-U\ Incylinder Pressure

Lfi/In-cylinder Temperature1 etc.

Physical Laws

Fuel ConsumptionEstimation

Q

Result

Total Fuel

Consumption

During DrivingCycle

Validation

Dynamic Test bench

Measurements

Figure 1-3: Overview of the central theme.


1.3 Objectives of the Thesis

Four objectives are pursued in this work. Three of them are related to the

concept of using secondary valves on a SI engine:

• fuel consumption reduction

• roughness reduction

• vehicle simulation during driving cycle with

a secondary valve equipped engine

The forth objective is to realize a set of tools that allow a convenient way of

analyzing and quantifying fuel benefits of a vehicle during a driving cycles.

Using simulations and measurements this work provides insight of how sec¬

ondary valves need to be designed regarding speed and flow discharge. Fur¬

thermore the timings and various modes of secondary valves operations are

analyzed with the aim of providing optimal parameter settings when

employed on an engine. Finally the interaction of engine and vehicle simula¬

tion as a whole to determine fuel consumption benefits of a vehicle during a

driving cycle is analyzed and verified. The used simulations are made avail¬

able for further use to evaluate new ideas regarding engine process and their

impact on total fuel consumption during a driving cycle.

1.4 Main Results and Own Contributions

Secondary valves are capable of operating in both the anticipated regimes -

reducing effective valve overlap to lessen the roughness of an engine with

aggressive camshafts at low loads and low speed, and using early effective

intake valve closure to implement throttleless load control.

Applying secondary valves by means of early intake valve closing on a SI

engine can reduce the fuel consumption by up to 10% neglecting the addi¬

tional energy consumption of the actuation devices. Because of severe temper¬

ature drop in the cylinder after the secondary valve closes - a consequence of

the expansion of the cylinder charge - the burn rate is deteriorated. Thus at

very low load or idling the benefits of secondary valves on the pressure loop

1.4 Main Results and Own Contributions 9

are counteracted by the negative effects of slow burning. This applies not onlyto the secondary valves principle but to any concept of early intake valve clos¬

ing and needs close attention when implemented. The inevitable presence of a

small parasitic volume between the two valves further deteriorates the effi¬

ciency of the secondary valve approach and thus engine efficiency, but this

effect remains minor compared to the overall advantages.

During a MVEG-95 driving cycle a total of 6.2% fuel saving is obtained,

again without accounting for the energy needed to actuate the secondaryvalves.

By using aggressive camshafts which were tuned to give maximum power at

high engine speed regardless of idle performance, engine power is augmented

by more than 15%.The idle quality of the engine with these camshafts is

unacceptably harsh, with approximately 50% misfire rate. However, by using

secondary valves to shorten the effective valve overlap engine roughness is

reduced by a factor of nearly 3 at 980 rpm to approximately the level found in

series production engines. This means that essentially the power range of the

engine, from minimum to maximum, is increased by about 15% with the same

engine displacement and roughness quality simply by preventing unwanted

exhaust gas backflow from occurring using secondary valves.

All experimental results are obtained using secondary valve prototypes not

suitable for series production engines. The necessary further development is

considered to be beyond the scope of this thesis.

1. As claimed by the manufacturer.

11

Chapter 2

Cylinder Charge Control

Internal combustion engines used in passenger cars operate over a broad

speed and torque range. Optimizing engine performance with respect to power

and emissions needs engine adjustment at every operating point; usuallyachieved exclusively with spark and fuel injection timing adjustments. These

are the only adjustable parameters that can be set independent on crankshaft

position. However, charge mixture preparation and the resulting burn rate, keyfactors for good engine performance, rely on more than just spark and fuel

injection timing. In order to obtain additional adjustment possibilities, other

devices have been proposed, e.g. intake runner length variation, intake port

disabling, swirl generation and variable valve timing. The latter is commonlyused with a fixed connection to the crankshaft, thus offering limited flexibility.Variable valve devices without this restriction open a new dimension of engine

control, making them worthy of detailed study.

2.1 Throttle Plate and Fixed Camshaft Timings

Camshaft timings are always a compromise between good cylinder filling at

high load and speed and roughness at part load and low speed. This trade off

cuts the peak engine performance by up to 15% by reducing charge filling

capabilities ([1], Chapter 6). As long as camshaft timings are fixed with

respect to crankshaft position, load regulation can only be achieved by reduc¬

ing the density of the inducted gas mixture or by changing the quality of the

fuel-air mixture (EGR, lean burning). Throttle plates are generally used, which

restrict the intake cross-sectional area and thus the flow. In this way significant

12 Chapter 2 Cylinder Charge Control

pumping losses are generated which can be as important as 30% of the indi¬

cated mean effective pressure as reported in [4]. This drawback can be avoided

if variable valves are used to control load by early or late intake valve closure

but in the case of early intake valve closing these benefits can also be achieved

using a secondary valve scheme.

However, it should be noted that one of the reasons why the combination of

throttle plate and fixed camshaft timings are still widely used are the low

costs, its mechanical robustness and convenient manner of achieving cylinder

charge control. Any variable valve device will have to be compared with this

benchmark.

2.2 Variable Valve Actuation

An ideal valve train for automotive engines would be one that satisfies all the

following requirements:

• variation of lift

• variation of timing and phasing• variation of valve transition duration

Neglecting the pressure drop of the intake system, the local cross-sectional

area determined by the valve and port geometry, and the valve lift are the only

parameters controlling how much charge is inducted into the cylinder. (Swirland charge composition also depend on these and other parameters). An ideal

device would allow any variation of valve cross-sectional area A with time.

AValve = Fit) (2.1)

This degree of freedom is theoretically feasible but hard to achieve with real

mechanical systems. In addition it is probably not necessary to achieve the

ideal device performance to obtain substantial improvements.

A rough overview of variable valve timing systems is summarized in

Figure 2-1. Variable valve actuation methods using camshafts as main compo¬

nents are not investigated further. The reason is that even though secondary

2.3 Overview of WTDevices 13

Variable Intake Valve Modes

Phase Shifting Lift Variation Closing/Opening Slope

Direct Actuation Not

Relying on Camshafts

Indirect Actuation

Relying on Camshafts

Electric Hydraulic Pneumatic

Figure 2-1: Actuation overview of variable valves.

valves are used in conjunction with conventional camshafts the range of possi¬ble usage is almost as broad as with variable valve actuation systems that

replace camshafts. This high degree of flexibility is hardly reached with sys¬

tems relying on camshafts, making them less interesting for comparisons with

a secondary valve system.

2.3 Overview of VVT Devices

Many variable valve actuation systems have been proposed in the literature,

generally grouped into electrical, pneumatic or hydraulic systems. A goodoverview is given in references [5], [16], [27] and [28]. Some interestingdevices of each category are briefly introduced.


2.3.1 Hydraulic: Active Valve System (Lotus)

This hydraulic system is used on SI engines up to 4000 rpm. Its oil flow-rate

can raise to 35 1/min and requires a pressure of about 210-10 Pa, leading to

an energy consumption of approximately 3 kW. It works principally like a

conventional hydraulic three valve system. This device allows a fuel reduction

of maximal 14% and if admitting additional 10% HC emissions, NOx could be

reduced by 83%. [6]

2.3.2 Hydraulic: Variable Valve System (Ford)

Another hydraulic system - it uses a sophisticated energy recovering system

to minimize losses. The timing and valve lift can be freely chosen, althoughsome bouncing effect deteriorate the shape of the valve lift curve. One of the

biggest benefits is the small dimension of the system: it can be build into the

cylinder head reducing its height to as low as 50 mm. Different control strate¬

gies were developed and by combining valve lift and closing time almost

every strategy can be deployed. [11][12][13][14]

2.3.3 Pneumatic: EVT System (Philips)

Using a piston and pressurized air at 6-10 Pa a very fast valve actuation

system was build able to operate even above 7000 rpm. Magnets were used to

keep the piston on which the valve is attached at both end-positions. The air

supply and the high current needed to hold the piston in one of its end positionwere externally supplied. The system is too bulky to be implemented on a four

valve engine, nevertheless on a two valve engine the fuel consumption could

be reduced by up to 4.4% and the engine torque increased by 15%. Emission

were lowered by 60%. [7] [8]

2.3.4 Electric: Variable Valve System (FEV)

The FEV system is a fully electromagnetic device with magnets on both end

and springs as driving forces. Due to electromagnetic properties the mass

spring system cannot be closed faster than approximately 3.5 ms and the lift

can not be varied unless another system is used to approach both end-posi¬tions. The system was modified to allow the implementation on a four valve SI

2.4 Overview ofSecondary Valves Similar Devices 15

engine. A sophisticated technique, managing the current for the magnets,

ensures that most of the energy in the system is converted back and forth from

electrical (capacitors) to potential energy (springs) going through kinetic

energy during transients of the valve. Overall energy losses are therefore

stated to be around 1 J per cycle and valve. With an operating range from idle

speed to 6000 rpm, fuel consumption was reduced by up to 15% at part load

and 25% at idle speed. The system is said to be ready to be used in production

engines. [9][10]

2.3.5 Electrical: Variable Valve System (Aura System Inc.)

This device is very similar to the FEV system and is available for productionline engines. Claimed fuel savings are in the order of 4 to 10% and emissions

reduction about 20%. The valve lift is 9.5 mm and the transition duration as

fast as 2.5 ms depending on the valve lift. On a four valve four cylinder engineless than 1 kW of power is used at full speed to activate all of the 16 valves.

[15]

2.4 Overview of Secondary Valves Similar Devices

Only a few devices similar to the secondary valves proposed in this thesis

have been reported in the literature.

In 1959 A. M. Kamel used a rotary valve to achieve early valve closing on a

four-stroke single cylinder engine [17]. The secondary valve consisted of a

tube with an orifice at the side and was mounted into a case connected to the

intake port of the cylinder. By synchronizing the rotary valve speed with the

camshaft speed but changing its phase the orifice passed over the intake port

so as to cut the intake flow before the intake stroke was finished. For this, a

phase shifting device was mounted in between the rotary valve and camshaft.

He reported 12% fuel savings at part load on an engine which can not be com¬

pared to those used today. Nevertheless he recognized the potential of second¬

ary valves and could demonstrate their benefits. Interestingly, this paper did

not mention problems associated to the parasitic volume.


The Kamel approach was explored further by N. Okansishi, I. Fakutani and

E. Watanabe in 1990 [18]. The test engine used the cylinder head of a four-

stroke cycle engine and the crankcase of a two-stroke cycle engine with,

instead of a tube as secondary valve, a disk with an orifice mounted in

between the intake port and runner. No experiments where carried out with a

fired engine. Their main target was the enhancement of the volumetric effi¬

ciency and comparisons were made with reed valves and secondary valves not

only in the intake runner but also on the outlet of the cylinder. The authors of

[18] claimed that if the valve cut-angles and timings of the rotary disc valves

at the inlet and outlet of the crankcase were set properly, the volumetric effi¬

ciency could be raised by 10 to 15% compared to using reed valves over the

entire engine speed range of 1000 to 7000 rpm. Again, parasitic volume was

not mentioned.

In 1992 another group started investigation into the secondary valve topic by

implementing a rotary disk valve on a supercharged four-stroke natural gas

engine [20]. Neither a disk nor a tube was used but a plate mounted on an axis,

rotating in a case, similar to a throttle plate but with no stop. The rotary valve

was driven by the camshaft with a toothed belt and could be phase shifted over

a range of 400° CA by help of a variable timing device using a helical gear

system. The final results were aimed at equal NOx emissions, thus not for

direct fuel comparison. Nevertheless for the first time the effects of the para¬

sitic volume were mentioned.

The same device was tested a year later on a four-stroke four cylinder gaso¬

line engine [21]. The secondary valves were driven synchronously by an AC

servomotor with the intake camshaft. The engine head was modified to a

higher compression ratio from 9.5:1 to 14.0:1. The effects of the parasitic vol¬

ume were not investigated any further but stated to be insignificant if the para¬

sitic volume was small enough. The simulations were done with an one-

dimensional model but only in motored conditions. Thus the deterioration of

the burn rate caused by the sharp pressure drop after the rotary valve closes

during the intake stroke was not analyzed. Even though the authors of [21]

expected bad burning to be the main reason for the observed lack of improve¬ments at very low load. In addition the rotary valve reopened again at low load

2.4 Overview ofSecondary Valves Similar Devices 17

because of its rotatory actuation, reducing the benefits. The results were

encouraging but no absolute values were given in respect of fuel consumption

savings.

Another approach to implementing secondary valves reported in [22] is to

use a shutter valve, a mechanical device sandwiched between engine inlet

manifold and inlet valve, consisting of two small plates with orifices sliding

on the top of each other. One is connected to the cam which is synchronizedwith the crankshaft. The other one to a special lever which acts as a load con¬

troller actuated by hand. Simulation and measurements were done and the

authors of [22] reported up to 30% rise in engine efficiency. Neither the prob¬lem of the parasitic volume nor the burn rate change due to early valve closingwas addressed.

19

Chapter 3

Secondary Valves

There are many different means of implementing the secondary valve con¬

cept to attain engine improvements. Depending of the implementation chosen,

the requirements become more or less severe.

In this section, starting by considering an ideal engine cycle with and with¬

out secondary valves, basic information is collated and the most relevant sys¬

tem requirements are found. The use of the secondary valve technique offers

some interesting opportunities to improve engine performance beyond chargeand roughness control.

3.1 Application of Secondary Valves

The main objectives of this thesis, part load optimization and the resultingfuel consumption savings are seen as most promising benefits of using sec¬

ondary valves. Roughness reduction with the help of secondary valves is an

additional benefit, especially if camshafts with large valve overlap are used.

3.1.1 Part Load Optimization

A conventional ideal engine cycle with constant-volume/pressure heat

release as shown in Figure 3-1 involves processes and assumptions listed in

Table 3-1. In this idealization valves have no transition duration from open to

close or vice versa.

20 Chapter 3 Secondary Valves

Throttle Plate Assumptions Secondary Valve

Process Points Points Process

1 Intake valve closes

Start of compression

1

Compression l->2 Isentropic l-»2 Compression

2 End of compressionStart of combustion

2

Combustion 2^3 Adiabatic

Isochoric

2^3 Combustion

3^4 Adiabatic

Isobaric

3-^4

4 End of combustion 4

Expansion 4^5 Isentropic 4-^5 Expansion

5 Exhaust valve opens

End of expansionStart of blow out

5

Blow Out 5^6 Adiabatic 5^6 Blow Out

6 End of blow out

Start of exhaust

6

Exhaust 6^7 Constant pressure 6^7 Exhaust

7 Exhaust valve closes

Intake valve opens

7

8' Start of intake 8

Intake 8'-»l Constant pressure 8^9 Intake

Secondary valve closes 9

Isentropic 9->l Expansion with

(vcy+vp) |

Table 3-1: Comparison of ideal processes.

3.1 Application ofSecondary Valves 21

Ideal early valve closing cycles have different low pressure parts comparedto conventional cycles, but the gas states during the pumping loop can be cal¬

culated with the same assumptions, summarized in Table 3-1. In case of earlyvalve closing, the time when the intake flow is cut off by the intake valve or

secondary valve determines how much work will be delivered by the engine.When conventional cycles and early valve closing cycles are compared, the

piston work during one cycle should be equal. Only then can fuel savings be

calculated correctly. It is also possible to determine the power generated bythe two cycles when consuming the same energy - i.e. investigating the extra

power generated from an early valve closing cycle from the same amount of

fuel. Both ways of comparison are interesting, but the former method is used

throughout this thesis.

To obtain same imep's in the secondary valve case as in the conventional

throttled case it is necessary to vary the point 9 in Figure 3-1. which is the

same as adjusting the closing time of the secondary valve.

logp A

logpatm

TDC

Parasitic Volume

BDC logV

Figure 3-1: Ideal engine cycle with and without secondary valve.


It is important to observe that even though computing the residual gas frac¬

tion in the throttled case is a delicate task, it requires an additional assumptionin the secondary valve case. In the conventional throttled case the assumptionis made that the residual gas in point 7 after expanding into the intake runner

is completely aspirated back into the cylinder. Thus the temperature in point 1

can be found by using the energy equation and the ideal gas equation. Havinga parasitic volume in the secondary valve case and postulating perfect instan¬

taneous mixing of residual and fuel-air mixture at any time of the cycle, some

amount of residual gas would remain in the parasitic volume after the intake

valve closes. Hence, at least one other system is needed to compute the gas

properties of the gas trapped in the parasitic volume. Moreover further

assumptions are necessary, as for example how residual gas mixes with the

manifold air when the secondary valve opens again. Only then the inducted

gas mixture of the next cycle is correctly modelled. By making the assumptionthat no residual gas at all remains in the parasitic volume after the intake valve

closes no other system is required. This assumption can be used for a first

order approach. Another simplification is made by neglecting some of the

energy stored in the pressure of the fuel-air mixture trapped in the parasiticvolume after the intake valve closed. This pressure is usually less than the

almost atmospheric pressure in the manifold. The energy stored in the para¬

sitic volume is partially recovered by aspirating fuel-air mixture through the

intake pipe as soon as the secondary valve opens again. This reduces the pis¬ton work during the next intake stroke. Certain modes of secondary valve

operation could partially recover this energy in another way, as demonstrated

in the next chapter.

The energy consumption of the secondary valve is not taken into account in

this analysis. The reason is that the realization of the secondary valve actuator

is not an issue of this thesis. Therefore only the effects of a secondary valve on

an engine process are analyzed and not the secondary valve device by itself.

The fuel savings as a function of parasitic volume and load and calculated

with the ideal engine process model are shown in Figure 3-2. Setting the para¬

sitic volume to zero corresponds to direct actuation of the intake valve, thus

obtaining best results with fuel savings of about 10% compared to the throt¬

tled engine. A realistic parasitic volume is roughly twice the clearance volume

if the cylinder head is not altered leading to an estimated 5% fuel savings.


0 20 40 60 80 100

Load [%]

Figure 3-2: Fuel savings as a function of parasitic volume (n times VTDç).

At full load no fuel savings are expected. The fuel saving difference of the

five different curves at 100% load comes from the fact that the secondaryvalve may be closed earlier than 180° CA ATC to achieve 99.9% of full load;the termination criterion of the secondary valve closing time iteration. There¬

fore in-cylinder pressure and temperature are not absolutely equal leading to

small losses or benefits in fuel savings.

The closing time of the secondary valve depending on load and parasitic vol¬

umes can be seen in Figure 3-3. At 100% load the closing time should be

180° CA ATC but as mentioned before, with a parasitic volume as large as 16

times the clearance volume, the secondary valve can be closed already at

130° CA ATC without changing mean effective pressure significantly. Fur¬

thermore it can be seen that a large parasitic volume does not allow any low

load. The reason is that even if the secondary valve closes before the intake

valve opens, a certain amount of fuel-air mixture is trapped in between both

1. Depending upon intake runner length and diameter.


180

160

140 -

<oo 120

Ü

< 100

CD

Ei- 80

D)

en60

o

Ü

>40

O)

20

0

! !

j^W

^%^/ Y

^0 l^^/ //2.i,< 8 16 = n

/ < /( / ""H

i

20 40 60

Load [%]

80 100

Figure 3-3: SV closing time as a function of load and parasitic volume.

valves. As soon as the intake valve opens the expansion process takes place

expanding fuel-air mixture from the parasitic volume into the cylinder. At

intake valve closing already too much fuel-air mixture is in the cylinder, pro¬

ducing higher load than desired. Hence with 16 times the parasitic volume it is

not possible to attain less than 60% of full load. The curve in Figure 3-4 shows

the maximal permitted parasitic volume if the secondary valves closes at

TDC.

It may be pointed out that with a very fast secondary valve it is possible to

reduce the load to zero indicated mean effective pressure even for large para¬

sitic volumes. This is achieved by letting residual gas penetrate the parasiticvolume until all fuel-air mixture is driven out. At this moment the secondaryvalve is shut. By this way, no fuel-air mixture is expanded into the cylinder

during the intake stroke and thus no load is produced.


0 20 40 60 80 100

Load [%]

Figure 3-4: Maximal permitted parasitic volume for a given load.

Finally it should be noted that the goal is always to reduce the parasitic vol¬

ume as much as possible to avoid losses. Without engine head modification

the volume is approximately less than two and a half times the clearance vol¬

ume. According to Figure 3-3 this is sufficient to control the load with second¬

ary valves in the same range as with a throttle.

3.1.2 Roughness Improvements

Random variation of cycle to cycle engine work is mainly due to the inho-

mogeneous mixing of residual gas, fuel, and air and the resulting change of

burn rate. If the variation exceeds a certain threshold, the comfort of the

engine is perceived as poor. Reducing the valve overlap of intake and exhaust

valve reduces the internal gas recirculation and improves the burning behavior

of the mixture. Secondary valves can reduce internal residual gas recirculation

in two ways. First, as a result of higher manifold pressure and second, by

actively preventing backflow when the intake valve opens - an effective reduc¬

tion of valve overlap.


At manifold pressures less than approximately half atmospheric, choked

flow will occur through the flow cross-sectional area opened by the intake

valve, meaning that a relatively large amount of residual gas flows back into

the intake runner. As long as the pressure in the cylinder remains higher than

exhaust pressure residual gas escapes into the exhaust pipes. Shortly after the

intake valve opens the in-cylinder pressure is lowered inasmuch that residual

gas of the exhaust pipe flows back into the cylinder. If the exhaust duration is

prolonged ending when the pressure in the cylinder drops below exhaust pres¬

sure, less amount of residual gas flows back and thus the next charge is

enhanced. Furthermore the amount of residual gas flowing from the cylinderinto the intake runner depends on the pressure difference over the intake valve.

In the next cycle this residual gas will be inducted into the cylinder before any

fuel-air mixture. The fraction of total residual gas can be as big as 28% [29].

With load control by secondary valves the manifold pressure remains almost

constant at slightly less than atmospheric conditions due to friction losses. The

pressure differences over the exhaust and intake valves are lowered reducingthe residual gas flow. Therefore, compared to conventional throttled engines,the time until the in-cylinder pressure drops below exhaust pressure during the

low pressure loop is reduced. Both effects result in a smaller residual gas mass

in the cylinder after intake valve closing and thus a better filling and burn.

These improvements will be achieved automatically if load is controlled with

secondary valves.

If the secondary valve is used for load and roughness control then it has to

remain closed after the intake valve opens. Only a small amount of residual

gas flows into the parasitic volume to equalize the pressure difference. Burned

gas is expelled from the cylinder through the exhaust port without any gas

flowing into the remaining intake runner. As soon as the intake stroke starts

and the in-cylinder pressure equals manifold pressure the secondary valve is

opened and normal fuel-air mixture induction takes place. The result is a low

amount of residual gas in the cylinder charge and thus a better burn, leading to

augmented engine power. This technique can also be realized with a reed

induction valve which blocks the flow section passively when the pressure

gradient is reversed - often used on two-stroke engine but also on four-stroke

engine. K. H. Menzel claimed fuel savings of 15% on a four-cycle engine and


Alfa Romeo measured torque improvements of 25% at 1000 rpm again on a

four-cycle engine with only 1% of flow friction at higher speed.

[23] [24] [25] [26]

The benefits of roughness control with secondary valves are the ability to

operate with large valve overlap, producing higher torque at full load and

speed and still gaining engine performance at part load by means of roughnessreduction.

3.1.3 Idle Speed Reduction

Idle speed is often a compromise between very low speed and important

roughness or higher speed and a more smoothly running engine. As long as

fuel consumption does not matter the latter is generally preferred. With strin¬

gent emission and fuel consumption restrictions the aim of reducing idle fuel

consumption has become important. As much as 296 seconds of the MVEG-

95 cycle totaling 1180 seconds consist of idling and typically 15% of the fuel

consumed is spent there.

With a secondary valve for roughness control, idle speed can be lowered to

as much as 600 rpm reducing substantially the idle fuel consumption.

3.1.4 Cylinder Cut Off

Cylinder cut-off is a technique in which instead of running all cylinders at

part load, one or more cylinder are switched off allowing the remaining cylin¬ders to run at higher load to produce the same overall torque. Hence, the fired

cylinders operate in a region of better efficiency. In a conventionally throttled

engine this can be done by stopping fuel injection. But the pumping work of

the non-working cylinder remains an additional loss which needs to be made

up by the remaining working cylinders, reducing the overall efficiency. With

secondary valves, cylinder cut-off is done likewise but only small pumpinglosses are generated because of the almost atmospheric manifold pressure.


Nevertheless, wall cooling of those cylinders that are cut-off has to be pre¬

vented, otherwise HC-emissions during the first couples of fired cycles are too

high. Cyclic cylinders cut-off with just one or two unfired cycles per cylinder

can prevent this effect.

3.2 Benefits and Drawbacks of Secondary Valves

Any system has its benefits and drawbacks. Both aspects are discussed in the

next chapters and are used to judge the overall performance of secondaryvalve devices.

3.2.1 Benefits

A secondary valve is considered to be used in conjunction with conventional

intake and exhaust valves. The advantage of this dual system is that if the sec¬

ondary valve fails to work on one cylinder this cylinder can be switched off by

disabling fuel injection. Although the other cylinders need then to run at

higher load, the engine at least continues to work.

Compared to intake or exhaust valves, the secondary valve is never in con¬

tact with the fuel air mixture during the burning phase. This allows the design¬ers to use plastic parts that withstand short peak temperatures of 800° K

occurring when residual gas flows back from the cylinder into the intake run¬

ner after the intake valve opens. Plastic valves are cheap to produce in any

arbitrary shape and can be of lightweight design.

In the conventional throttled engine case the manifold filling delays any

engine response to throttle changes. In contrast, in the secondary valve con¬

trolled engine the load change can be achieved in the very next intake stroke

of every cylinder by adjusting the secondary valve timing.

In practise there is always an inequality in torque produced by the cylinders,

partly due to cycle to cycle variation but also because of small intake flow dif¬

ferences caused by runner length or cylinder wall temperature variations. If an

in-cylinder pressure sensor is present, these variations can be eliminated by a

feed back control system acting on the secondary valves.

3.2 Benefits and Drawbacks ofSecondary Valves 29

3.2.2 Drawbacks

It is evident that the use of secondary valves requires more moving parts.

And with more parts the probability of mechanical failure increases.

The fuel consumption reduction depends highly on the engine size. A small

engine almost never operates at very low part load conditions and so pumping

loop losses are not as significant compared to those obtained from large

engines.

How the secondary valve may be powered and where the energy supplycomes from has not been investigated in this work since these aspects dependon the chosen actuator design. The overall benefits are strongly dependent on

how the external energy supply is implemented and what kind of actuation

would satisfy the demand with as little as possible energy consumption.

For a secondary valve, the transition duration is important for efficiency and

thus also for the engines overall efficiency. If the valve acts as slow as the

intake valve, double flow losses occurs lowering the intake pressure and

increasing the pumping work. Especially at part load the closing speed should

be less than 3 ms whereas at higher speed a throttle may become necessary to

control the load.

Probably one of the most important problems with any early intake valve

strategy is the sudden pressure drop in the cylinder after the intake valve

closes. This might cause condensation of fuel during the expansion after earlyvalve closing. Moreover the decrease of turbulence caused by the increased

duration from early intake valve closing until ignition is another drawback.

Both effects prolong the burn duration and reduce the overall efficiency. The

lower the load the lower the temperature and the worse the burn rate leadingalso to higher HC emissions.

E -e**^ r I

^ät^ri

31

Chapter 4

Models of Engines with

Secondary Valves

The ideal engine cycle calculation shown in Chapter 3 is not able to accu¬

rately estimate secondary valve timings, fuel savings, backflow phenomena or

burn rates and thus a better model will have to be used. An important issue in

any overall engine model is the balance in complexity and detail amongst the

process sub-models. At present, it is not possible to construct models that pre¬

dict engine operation from the basic governing equations alone. Therefore the

objectives of any model development effort should be clearly defined, and the

structure and detailed content of the model should be appropriate to its objec¬tives.

4.1 Engine Model Requirements

As explained in Chapter 3.1, part load optimization is the major issue of the

simulation performed in this work. Thus the characteristics of the engine

model will focus mainly on this topic.

Load Control Modes - Having two different engine configurations, one with

a throttle plate and the other with secondary valves, it is necessary to either

design a model able to switch between the two modes of charge control or to

have two different models. If the former approach is chosen, any sub-model

has to be suitable in both cases. Since the throttle plate and the secondaryvalve lie in front of the cylinder, care must be taken to ensure correct manifold

32 Chapter 4 Models ofEngines with Secondary Valves

gas state. In the secondary valve case the load depends also from the gas state

in the parasitic volume which therefore needs to be tracked during the entire

engine cycle. Likewise the manifold gas has to be correctly conditioned for

the next cycle.

In-cylinder Pressure - There are three reasons why this pressure is one of the

key variables of an engine model. First, the calculated pressure by itself will

allow the validation of the model with measured in-cylinder pressure traces.

Second, in the secondary valve case the pressure drop caused by the expansion

during the intake stroke allows some conclusion about the burn rate behavior.

Thirdly, calculation of indicated mean effective pressure needs accurate pres¬

sure computation. Only when both imep's, throttled and with secondary valve,

are equal it will be possible to compare other values and to draw correct con¬

clusions.

In-cylinder Temperature - The estimation of a burn rate depends strongly on

the gas temperature in the cylinder. Backflow and thus temperature in the par¬

asitic volume decides upon the requirements of temperature resistance of the

secondary valve. Pressure and temperature gives a good base to qualifyobtained results.

Residual Gas - Engine roughness is mainly due to the amount of residual gas

in the cylinder charge. Therefore it is necessary to precisely calculate the

residual gas fraction in the cylinder charge as well as in the intake runner.

Moreover the temperature of the fuel-air mixture, inducted during the intake

stroke, also strongly depends on this variable.

Burn rate - Burn rate prediction is eminent. The influence of early valve

closing on charge temperature and therefore fuel-air mixture quality is very

strong. Without adaptation of the burn rate, the impact of early valve closingon the thermodynamic process will not be reflected correctly, leading to too

optimistic fuel consumption benefits.

Variations ofSecondary Valve Properties - Position and speed of the second¬

ary valve have an important influence on the engine system efficiency. A slow

valve will act more like a butterfly valve, i.e. throttling the engine and causingadditional pressure losses. Mounting the device far away of the intake valve

4.1 Engine Model Requirements 33

results in a considerable parasitic volume reducing the efficiency of the sec¬

ondary valve dramatically. To analyze these phenomena the position and

speed of the secondary valve will be varied in the simulations. This impliesthat the runner is split into two parts: one in front of and the other behind the

secondary valve; the latter representing the parasitic volume. Two systems are

necessary since only then it is guaranteed that heat losses in the intake pipeand its consequences can be compared independently of the secondary valve

position.

Pressure Waves in the Intake and Exhaust Pipe- Pressure waves have a sub¬

stantial impact on the cylinder charge at intake valve closing. Good tuning of

intake and exhaust pipes increases the volumetric efficiency of the engine sig¬

nificantly. Therefore an engine model that does not include pressure waves

dynamics can hardly deliver accurate cylinder charge prediction.

Neglected Effects- Some engine models allow prediction of knock. This is

important if high load performance has to be calculated. When computing

engine cycles the spark timings need to be set preliminarily or iterated for

maximum break torque. In the latter case it is essential to include a knock pre¬

diction sub-model in the engine simulation. Otherwise the engine model

would allow spark timings at which a real engine would heavily knock. Since

secondary valves are deployed solely at part load, knock prediction is not a

requirement of the engine model.

Inasmuch as emission predictions requires a complete model of the chemical

reactions during the fuel-air mixture burning, the computation complexityincreases dramatically. This work is a first approach of evaluating secondary

valves on the engine process and therefore emission predictions will not be

addressed. This allows a much simpler model to be draft without claiming that

the neglected effects should not be considered in further investigations.

The requirements are summarized in Table 4-1.


Requirements Important Not important

Conventional load control (throttle) /

Secondary valve load control /

Fuel consumption predictions /

In-cylinder pressure predictions /

In-cylinder temperature predictions /

Residual gas predictions y

Burn rate predictions s

Variation of the parasitic volume /

Variation of SV transition duration /

Analysis of pressure waves »a

Emission predictions /

Knocking predictions /

a. Pressure waves are taken into account by boundary conditions of the simulation.

Table 4-1: Requirements for the model.

4.2 Selection of an Engine Model

Engine models can be classified roughly into fluid-dynamic or thermody¬namic approaches. Whereas fluid-dynamic engine models are multidimen¬

sional and allow geometrical engine information to be included,

thermodynamic models are zero- or quasi-dimensional. The introduced

requirements include considering pressure waves, which need fluid-dynamicmodelling techniques. However, if measured pressure traces at intake and

exhaust engine ports are available and used as boundary conditions for a zero-

dimensional engine model, then a thermodynamic model will be sufficient to

satisfy the mentioned requirements.

4.2 Selection ofan Engine Model 35

Thermodynamic energy related models are often classified into single zone

or multi zone models. The main difference between one and multi zone model

is the more accurate prediction of in-cylinder temperature. The reason is that

information on flame propagation, spray shape or cylinder wall effects can be

included into the model. Accurate temperature calculation in the combustion

chamber is important for improved emissions prediction. Since neither the

temperature field in the cylinder nor emissions are a primary focus of this

work there is no necessity to implement a multi zone model.

The in-cylinder pressure is very similar for all thermodynamic models, since

pressure equalization happens with sonic speed across every zone in contrast

to energy transport in form of mass or heat flux.

Burn rate estimation from chemical reaction of the fuel with the air is an

extremely complex process often done for knock detection or emissions pre¬

diction. In the vast majority of cases semi empirical models are used to deliver

sufficiently accurate results. By using such a model with at least pressure and

temperature as parameters, the impact of early valve closing on the burn rate

can be estimated.

In any case an engine model requires the computation of the properties of

state of two gases: fuel-air mixture and burned gas. By knowing the mass frac¬

tion of the burned gas in the total considered gas mass, every property of the

three gas mixtures, fuel-air, residuals and burned fuel-air are derived. Only the

consideration of all three gas mixtures in every system leads to accurate tem¬

perature and pressure predictions.

Considering all the above remarks, it can be concluded that a single zone,

two components model (three different gas mixtures) with three systems, one

in front of the secondary valve, one behind and one being the cylinder itself,will satisfy all requirements. Such a model will be able to simulate the impactof the secondary valve on the engine cycle process and delivers all results

needed to analyze the questions raised in the previous chapters.


4.3 Model of the Secondary Valve and the Engine

Regardless on any sub-model or equation used, the organization of the pro¬

gram remains the same as summarized in the flow-chart of Figure 4-1. The

initial state of the system is unknown and must be estimated. Afterwards the

cycle is iterated until the state variables of the system at cycle end, that is end

of the intake stroke, matches the state variable at cycle start which is the

beginning of the compression stroke. The closing time of the secondary valve

is estimated from the indicated mean effective pressure of the throttled cycleand the ideal engine cycle computation of Chapter 3. A double iteration is

needed in the secondary valve case: one iteration to satisfy cyclic conditions

and the other to adjust the load. Finally, when the imep difference of the sec¬

ondary valve equipped engine and conventional engine is sufficiently small,

fuel consumption benefits, residual gas fraction etc. can be compared and ana¬

lyzed.

The model is written in Matlab/Simulink allowing simple modular structure

of the code. Each physical model is represented by a sub-block worked out as

a S-function and incorporated into the main core simulation block. The simu¬

lation starts at intake valve closing and ends 720° CA later, after which a new

cycle is initialized and integration is restarted.

The engine model consists of three systems as shown in Figure 4-2, con¬

nected to each other by mass and energy relations.

4.3.1 Main Equations

Before the mass and energy equation are presented for each sub-system, the

properties of gas mixtures involved in the process are analyzed. Two types of

gases are distinguished: fresh fuel-air mixture and burned fuel-air mixture.

Without modelling the fuel injection and its evaporation, a stoichiometric air

to fuel ratio of the fuel-air gas is assumed in the manifold. In the case that

these two gases are combined a mass ratio 2, is introduced,

% =-^5£d

(4.1)mtotal


Initialize core parameters

t

Preprocessing of crank angle depending values

Not reached

Not reached

Not reached

I-•h Run engine simulation

Set SV closing time

ïRun engine simulation

Reached

Final results

c

"5b,

c

<©.,

O

CDC

"enc

LU

TDCDQ.Q-

'=3O"

LU

CD

_>

COT3

CoÜCD

CO

Figure 4-1: Chart flow of the simulation.


System Description

01 Volume of the intake runner in front of the secondaryvalve.

12 Volume of the intake runner after the secondary valve,

referred to as parasitic volume.

23 Cylinder with boundary cylinder wall, piston, intake and

exhaust valve.

Figure 4-2: Model of the engine with secondary valve.

allowing the calculation of the new properties with help of the two basic gas

properties.

Ideally the compression stroke starts with only fresh fuel-air mixture in the

cylinder. A burn rate describes the heat release of the fuel during burning and

thus the conversion from fresh fuel-air mixture to burned gas. This burned gas

is expelled during the exhaust stroke and fresh gas enters the cylinder duringthe next intake stroke. In reality burning is incomplete leaving some unburned

fuel-air mixture left when the burning has ended. Moreover the cylinder can


never be totally scavenged; fresh fuel-air mixture always mixes with burned

gas. With back flow phenomena residual gas even enters the intake runner.

The conclusion is therefore, that two gases have to be considered in every sub¬

system:

• residual gas r : a mixture of unburned and burned fuel-air gas

• air mixture am : fuel-air mixture only

Finally all mass equations for the systems and their correspondent gas mix¬

tures can be expressed as shown in (4.2) with the indices: am for fresh fuel-air,

r for residual gas mixture, ab for the air mixture converted to burned gas, at

for the total air equaling the sum of air in both gas mixtures # (one number)

for the cross-section and ## (two numbers) for the system itself.

dmamOldQ

dmrOl

dQ

dmam\2

dQ

dmrl2

de

dmami?)

dQ

dmr23

mamO "*" maml

CO

mr$ + mri

CO

~ mam\ "*" mam.2

CO

-mri + mri

CO

~mam2

CO

- mr2 + mr2

(4.2)

dQ co

dmat23 -mam2-(l-^)mr2 + (l-^)mr3"de-

=

œW

dmbt23 ~ \T1 + £>r3dQ co

+ m. bxB


The first law of thermodynamics for a open system, the energy equation, can

be written as

È = Qw-W+^mjhj (4.3)

j

with

È = jt(mu) (4.4)

denoting the total energy change in the system, Qw is the total heat transfer

into the system across the system boundary and W the work transfer rate

equaling pdV.

By using ideal gas relations and changing the derivative from time to crank¬

shaft angle the energy equation leads to the temperature differential equationof every system.

Starting with system 01 (see Figure 4-2 on page 38), only two mass flows

are possible: across sections 0 and 1. Heat transfer takes place from the gas to

the wall or vice versa and the lack of any moving boundary eliminates the

work transfer term.

dTQl_

<201 + HamO + Ham\ + Hrp + Hr\

dQ~

<D(cvam(r0l),nam01+cvr(r0l)mrOl)

dmam0l dmr0l (4'5)

"rfë-mMam01 + -rf8-">MH)l

^Sam^OlKmOl + Sr^OlKoi)

Enthalpy flow is denoted as H, heat flux as Q, specific internal energy as u

and specific heat at constant volume as c.


The signs are chosen to be positive for flows going into a system. At the con¬

nection of two systems the signs for same flow variables are positive for one

system and negative for the other.

System 12 is similar to system 01 :

dT\2_

Ql2-Ham\-Hr\+ Ham2 + Hr2

dQ~

a(cva(T12)mal2 + cvr(T12)mrl2)

dmam\2 dmr\2 (4"6)

^Cva^T\^ma\2 + Cv^Tl^mr\2)

The third system 23 requires two additional terms which describe the burn¬

ing heat release QB and the power Wp delivered to the piston.

dT.23

_

Ö23 + QB ~ Ham2 - Hr2 + Hr3 + Wp

dQ ^cvafr23>mat23 + Cvbt^T23^mtb23 )

dmat23 dmbt23 (4'7)

——COM^.oo + ——COM» .Oo

dQat23 dQ bt23

^cvafT23^mat23 + Cvbt(?23>mtb23">

With this set of equations the engine model satisfies the requirements listed

in the last section.

4.3.2 Secondary Valve Characteristics

Implementing a secondary valve means computing a flow through a cross-

sectional area from wide open to fully closed as done for poppet valves.

Instead of precisely modelling the dynamics of the secondary valve, a time

dependent function was used describing the transition duration. Every closing

profile can therefore be implemented which is useful for the comparison of

different mechanical systems. For the final computations, a constantly

decreasing cross-section was used. The slope of the cross-sectional area of the


Secondary Valve

Flow Restriction Area

A SV Open End SV Close Start

SV Open /*-

SV Close

_

Transition Duration

Transition Du ration_

SV Open Start SV Close End

Crank Angle

Figure 4-3: Secondary valve lift.

used pneumatically actuated secondary valve shows a more parabolic shape.Nevertheless, a linear approach was implemented resulting in an inaccuracy of

less than half a millisecond of the secondary valve transition duration.

Because the secondary valve transition duration relates to time but the simu¬

lation is based on crank angle a time to crank angle transformation is needed,

shown in Figure 4-4. The higher the engine speed the less effective is the sec¬

ondary valve, i.e. it will act more like a throttle than a binary device. Moreover

the high frequency of the secondary valve operation at higher speed is hard to

realize. This is one of the reasons why at high engine speed a throttle should

be used rather than secondary valves.

4.3.3 Gas Components and their Properties

Gas properties are one of the most important values in an engine model.

Since residual gas is a mixture of burned and unburned gas, its gas propertiescan be calculated as weighted averages of the properties of the two basic

gases.

High temperature effects are considered for the burned fuel-air mixture tak¬

ing dissociation into account. This is not done for the fuel-air mixture since its

importance diminishes with increasing temperature because fuel-air mixture is

converted into burned gas, reducing the total fuel-air mass. Components of the


4500

4000

3500

E

Ê3000TJCD

a. 2500<n

§,2000c

LU

1500

1000

500

6.25 ms

-i r

: 6.25 ms

IVC SVO SVC IVC

-200 TDC 200 400

Crank Angle [deg]

600 800

B SV Transition Duration

Figure 4-4: Transition duration of the SV in °CA depending on engine speed.

fuel-air mixture are fuel, 02, N2, C02, H20, CO and H2 with masses depend¬ing on the fuel air mixture ratio. In the case of the burned mixture the same

components are present. Zacharias [31] proposed an overall function describ¬

ing enthalpy, specific heat, and internal energy for the burned mixture depend¬

ing on temperature and pressure. This relationship is finally implemented,whereas for the fuel-air mixture the thermodynamic properties of each speciesare calculated and appropriately summed to get the overall thermodynamic

property of the fuel-air mixture. [30]([1], Chapter 4)

4.3.4 Burn Rate Estimation

One of the key modules of the simulation is the estimation of the burn rate.

The burn rate is recalculated after every engine cycle iteration which impliesthat the heat release will be changed for the next cycle. This alters the engineefficiency continually. In conclusion it is indispensable to consider burn rate

alterations due to early valve closing for reliable engine efficiency predictions.


Rassweiler and Withrow [35] proposed a simple and robust method of esti¬

mating the fuel-air mixture conversion from pressure traces. Having pressure

traces for the whole conventional engine operating area, the burn rate can be

extracted and implemented into the engine model.

For the engine with secondary valves, only a few operating points were mea¬

sured and therefore this method could be deployed only partially. Hence, a

function is necessary to describe the burn rate of the conventional engine with

secondary valves based on the state variables obtained during the engine cycle

computation.

Mass fraction burned versus crank angle profiles determined from first law

analysis of pressure data have an essentially universal dimensional shape. It is

therefore significantly easier to use a function describing this shape than esti¬

mate the burn rate from physical derived complicated equations without loos¬

ing significant accuracy.

Vibe [36] used an exponential function for the mass fraction burned.

xh = 1 - exp -a

Q-QQ\m+\.(4.8)

where 6 is the crank angle, 6q is the start of combustion, À9 is the total com¬

bustion duration and a as well as m are adjustable parameters.

Csallner [34] suggested using simple weighting functions of the Vibe param¬

eters starting from a reference engine operating point to compute the burn rate

at other engine operating points. By using fuel-air ratio, in-cylinder tempera¬

ture and pressure at a given crank angle, residual gas fraction and finally the

ignition crank angle, new values for the Vibe parameters À0, 0q and m can

be estimated with this approach. The implementation in the case of the con¬

ventional or the secondary valve equipped engine is shown in Figure 4-5.


Pa

Pressure Trace (Measured)

xb

^CA I

* 1

Rassweiler & Withrow

*XA Ixbi T

Vibe Engine with throttle plate

xbi

DCA ICsallner — Vibe Engine with secondary valves

°CA

Figure 4-5: Computation of burn rate.

The set of equations Csallner proposed for the burn rate estimation is

9 = 0 .+ 8.-8. V-+A8 ,

s s, ref i i, ref s, ref I/r (4.9)

e„ = vA(w-n*7 (4.10)

m - m nref ilhj (4.11)

9 is the crank angle and the index s referring to the start of the combustion, e

to the end and i to the ignition delay.


Equations (4.12) to (4.17) describe the parameters used in (4.9) to (4.11).

Considering that the conventional cycle is used as starting point for the esti¬

mation of the secondary valve burn rate, not all of the parameters are

unknown. Engine speed remains the same in both cases and by postulatingstoichiometric fuel-air mixture two of the six parameters can be predeter¬mined. Since in the secondary valve case the burn rate will be delayed, the

ignition timing could be set earlier but to guarantee the computation to be

conservative, the ignition was not altered. The remaining relationships used in

the engine simulation are shown in Figure 4-6.

Temperature and pressure drop at compression increase burn duration while

the reduction of residual gas has the contrary effect. In the secondary valve

case the burning start is delayed and the burn rate duration is increased which

leads to a shift of 1 to 15° CA in the 50% conversion point on the burn rate

curve. [53] [54]

T/Tref

P/P.ref

x/xref

1.05

0.95

a>

Figure 4-6: Weighting terms for the burn rate estimation.


Ignition (4.12)

430 - 9;/; =

430-<P,,re/

«*=1

*,= 1

Fuel-air mixture (4.13)

2

4 =

«i

2.2A, -3.74X. + 2.54

2.2^- 3.74X„/+ 2.54

2il -3.4X + 2.4

2^/-3-4W+2-4*l=1

Temperature

/r = 2.16ref 300

r

(4.14)

-1.16

Residual gas fraction (4.15)

xf = 0.088 —+ 0.912X v

300x

ref

T

gT = 1.33- ^/,300-0.33r.300

/Zy = 1

gr = 0.237

*x=»

X

+ 0.763xref

Pressure (4.16)

fP =

SP=

f r> ^a47^300

Pref, 300,

( n V0-28

J7ref, 300y

*, = i

Speed

f =J n

ö ri

(4.17)

1 + 400Az - 8xlQ5/n2

1 + 400/nrg/- 8xl05/n^71.33-660/n

n

1.33-660/nrg/

=

750/n + 0.625

n

75Q/nref+ 0.625


4.3.5 Wall Heat Transfer

Woschni [39] suggests a simple, but nevertheless sufficiently accurate equa¬

tion, which describes the average heat transfer from the cylinder gas to the

cylinder wall. This equation is based on Nussel and Reynolds term correlation

[38] of the dimensional analysis of convective heat transfer. Hohenberg [32]

slightly modified this equation to get better results. More sophisticated meth¬

ods often use local instead of average heat flux making the model in the con¬

text of this work unnecessarily complicated. For the sake of simplicityWoschni's approach is used.

Qw = hcAw(Tc-Tw) (4.18)

ir--Dm~^ m m 0.75-1.62m

,a in\h = CB p w T (4.19)

With the cylinder bore B taken as the characteristic length, w as a local aver¬

age gas velocity in the cylinder, p = pRT and C set to 130.

w =

V,.C-,S„ + Cn——(p — p )

1 p ln yyr ? m>

^r r

(4.20)

For the gas exchange period: Cj = 6.18; C2 = 0

For the compression period: Cj = 2.28; C2 = 0

_3For the combustion and expansion period: Cj = 2.28; C2 = 3.24-10

The exponent m in (4.19) is set to 0.8 and all units are in Si-units except the

pressure which is in kPa.

Heat transfer in the intake runner is modelled using equations derived to

describe heat transfer of a turbulent gas flow in tubes ([1], Chapter 21)

[33][48].


4.3.6 Spray Fuel Evaporation and other Considerations

It should be noted that after the secondary valve closed, fuel condensation

can become a severe problem. Condensation is a complicated process depend¬

ing not only on time but also on the particles in the fuel-air mixture. Simplefuel evaporation charts show that at atmospheric pressure condensation starts

merely at 288° K. Considering also the evaporation energy the air cools down

by about 20° K. Hence, the critical temperature of the air alone is 308° K [40].

Even in the conventional throttled case, some amount of fuel is not evapo¬

rated, not burned or escapes from the cylinder through some leakages. This

amount was considered globally to be 3.5% of the total fuel mass [40], [55]. In

the secondary valve case this amount was increased to 5%. Together with the

burn rate estimation of Csallner in Chapter 4.3.4, the heat release is modeled

in a conservative way. The good correlation with the measurements justifiesthe method used for burn rate prediction (see next section).

4.4 Validation of the Simulation

Having a conventional and a secondary valve model, two validations are nec¬

essary. Both were done with measurements on a series production four valves

two liter engine supplied by Ford Motor Company.

4.4.1 Conventional Engine

Two representative engine operating points are shown from the overall 80

points calculated. The first point at 1500 rpm and 4.8-10 Pa imep and the

second at 2000 rpm and 1.9 TO Pa imep. The results and correlations with

the measurements are given in Table 4-2.

imep .j— imep

^thr,sim ^meas irir. sA^-t\

lmeP'error, rel.(meas,thr)= ~~

l0° (4-2l>

un*Pmeas


imePsv,sim ~ lmePthr,simim ,,

_

lmePerror,rel.(thr,sv)= JZ~ 10° <4'22)

imerthr,sim

fuel masserrorrel{meaStthr)= (4.23)

fuel massthr>sim-fuel massmeas

fuel mass

100

meas

fuel mass saving = (4.24)

fuel mass 100 — imep, ,tl x

J sv,simyerror,rel.(thr,sv) -. *-.

fuel mass., .

Jthr,sim

total error , , ., N= imep

, , +? \—(4.25)

rel.(meas,thr)Ferror,rel.(meas,thr)

v y

fuel mass -, f .-, ^J error, rel. (meas, thr)

Regarding all of the 80 operating points analyzed, errors of indicated mean

effective pressure were always in the range of ±4% and errors of fuel con¬

sumption, after linear correction with the imep, never exceeded ±5 %. The

total error is defined as the sum of the indicated mean effective pressure error

and the fuel consumption error. None of the 80 calculated points showed a

total error of more than ±5 %.

The effect of the burn delay caused by early valve closing on the in-cylinder

pressure can be seen in Figure 4-8 or on the in-cylinder temperature in

Figure 4-10. The fuel consumption saving are reduced to only 3.5%.


n 1500 2000 rpm

imepmeas 474631 192120 Pa

imepthl.sim 482813 196446 Pa

imePerror,rel.(meas,thr) 1.7 2.6 %

imePsv,sim 483002 188980 Pa

lmePerror,rel.(thr,sv) <0.1 -3.8 %

fuel massmeas 1.8010-5 9.09 -10~6kg

cylinder • cycle

fuel massthr?sim 1.85-10"5 9.05 10"6kg

cylinder cycle

tuel rnasseiror)re]_(meaSjtjjr) 2.8 -0.4 %

total errorrel(measthr) -1.1 3 %

fuel masssv sjm 1.71-10"5 8.42-10"6kg

cylinder cycle

fuel mass savingcorr -7.6 -3.5 %

burn rate center shift +3.6 +9.3 °CA

residual gas reductionrel,(thr,sv) -10.7 -15.7 %

SV closing timesim 418.6 377.3 °CA

Table 4-2: Example of simulation results.


10

o106W(O

Q_

<DT3

|io5

10

:::::::: : : c : : : : i :.:;:::: : y- - --

yu~

y-

JLJL-^L-l-

ii nullit;

j u ^ i

SV

- - Measurement

i i i i , |1 1 ' . i i

;.

'- ^^-*~S'!.' i iii i i

;

-

'

-VV^-

! ! ^""Sa !'

- _

i„ - -

'-

^t '

- -

' L >_

i L.

-----|--- r-

r

V* -i-i

>£sl:;- : : : :-

J*Sj. 1 h , +

r

____n___-

T XNr ">- - -^V - -

r-

-,- -

; ; r xs,"

: x*;-^-,.1

-..

XV u.\ -

'

• ',',', ,' XX'

'\'

J- - |- -^V'^r^»^»-^<<.«^'i^''^a-^*-^S^-

1 1

i i

1 1 , 1 , 1 1 , , 1 1

, , 1 1 , , ,

! ! \ , , ; , i,.

1 1

, , ,

iii i iii

1

10-5 le"4 ID"3

Cylinder Volume [rri\]

Figure 4-7: In-cylinder pressure at 1500 rpm and 4.8 bar imep.

107

CO

^i

£ 10

wwCD

0)T3

|io5

10

: ::: j:: : ::::-::I ' 1

i i 1

..„,,— | nfouie_j L

SV

- - Measurement

, i i J. J i-

1 1

, , , i i 1 i iii i l

• j*~

iii i iii

&tt*^, < 1 1 1 1

j j- ,- l H

" " ~-' "!jSb- "' ' <---'--

_i

_

-

r tJ-" -

7 "I~ ! =—"-^Sv7 -_1---w|---r---|-~

"

"r (S:

"

X:X n" ~

Xj~- -

r-

-i- -

, , i\S, , i >t i

: : i x'*^—;A,-,—

-, -\ - -

--•r-V -

t -! "^X^"*' ""'

" " ~ !"

-

-, --r-r\

^^p»f^:-

T_ _

r.

,

' '

iii i iii

1

,-510

"

10~*

10

Cylinder Volume [m3]

Figure 4-8: In-cylinder pressure at 2000 rpm and 1.9 bar imep.


2500

^2000o

CD&_

2 1500CDQ.

CD

CD

C

o

1000

£ 500

0

I1

.- .. Tr THRCy

Tn SVCy

I

1 \

1 '\

1

j.

1

-200 0 200 400TDC Crank Angle [deg]

600 800

Figure 4-9: In-cylinder temperature at 1500 rpm and 4.8 bar imep.

2500

^2000

CD

S! 1500CDQ.

ECD

CDac

">>Ü

1000

S 500-

0-200 0 200 400

TDC Crank Angle [deg]

I

CyTn SVCy

'

600 800

Figure 4-10: In-cylinder temperature at 2000 rpm and 1.9 bar imep.


4.4.2 Engine Equipped with Secondary Valves

The secondary valve engine model was validated with the same Ford Zetec

4-cylinder 2.0-liter engine as used to measure the part load engine map.

Mounted on every intake port was a secondary valve (Chapter 7.1.2) and the

in-cylinder pressure was measured once with secondary valve load control and

once with throttle plate load control. Both measurements could be accurately

reproduced by the simulation. The closing point of the secondary valve, which

was set for MBT during the measurement, was calculated by the simulation to

close at almost the same crankshaft angle. The following two figures show

graphically the correspondence achieved - Table 4-3 shows the main results.

imePmeas 301773 Pa

ePthr,sim 298270 Pa

imeperror rej (meas,thr) -1.2 %

imePsv,sim 305805 Pa

lmePerror,rel.(thr,sv) 2.5 %

fuel massmeas 1.3110"5 (kg)/(cylinder cycle)

fuel massthrsim 1.28-10-5 (kg)/(cylinder cycle)

tuel rnasserrorrei.(meas,thr) -2.3 %

fuel massSV)Sim 1.28-10-5 (kg)/(cylinder cycle)

fuel mass savingcorr -2.6 %

burn rate center shift +1.2 °CA

residual gas reduction^(thrsv) -53.8 %

total errorrel(measthr) -3.5 %

SV closing timeSjm 373.15 °CA

SV closing timeExperiment 375 °CA

Table 4-3: Secondary valve validation at 983 rpm.

4 4 Validation of the Simulation

10

CO

Q_

g 106co

COCD

CD

|io5

10

Throttle

Measurement

io"5 io-4 io-3

Cylinder Volume [m3]

Figure 4-11: In-cylinder pressure at 983 rpm and 3 bar imep - conventional.

j10

cc

cl

o 10É

COCD

0T3

|io5

10

SV

Measurement

10 104

10-3

Cylinder Volume [nr]

Figure 4-12: In-cylinder pressure at 983 rpm and 3 bar imep - with SV.

"£. Li »i'JÖ^

57

Chapter 5

Main Results of the

Simulations

80 different engine operating points were simulated, corresponding to the

engine fuel map measured under part load operating conditions

(Chapter 7.2.2). The speed range was from 800 to 4000 rpm and the load

ranged between zero and approximately 4 10 Pa bmep. The goal of the sim¬

ulation is the computation of the following maps of an engine equipped with

secondary valves:

• fuel savings map• residual gas fraction reduction map• burn delay map• secondary valve timing map

The first three maps are compared to their counterparts of a conventionallythrottled engine. The fuel saving map is of primary importance, since this

work focuses on fuel consumption. Moreover, fuel maps will be used in

Chapter 8 as main input for the driving cycle simulation. The secondary valve

timing map is the base for any control schemes and needed when further

investigations are made. For example, a feed forward controller may use the

information given in the secondary valve timing map to convert the gas pedal

position into an electronic signal, triggering the secondary valve actuator. The

burn delay map as well as the residual gas map show the trend of this particu¬lar engine behavior over the entire operating range, emphasizing drawbacks

regarding efficiency and benefits regarding comfort. A roughness map would

58 Chapter 5 Main Results of the Simulations

be convenient but cannot be obtained directly from the simulation since no

engine dynamics are considered in the engine model. However, inasmuch as

residual gas influences the magnitude of the engine performance directly, it

remains a good measure for comfort quality.

5.1 Analysis with Willan's Approach

The quality of the measurements and of the corresponding simulation results

can be analyzed using the Willan's approach as explained in Chapter 1. The

region of interest in the fuel consumption map is shown in Figure 5-1. All

measurement as well as the corresponding engine cycle simulations were con¬

ducted in this region. Following the Willan's approach, a graph of bmep versus

fuel mep can be drawn at each constant engine speed cross-section of the fuel

map. The linear equation describing the relation between bmep one.fuel mep is

given by

bmep - e -fuel mep + (fmep^ + imeppl 0) (5.1)

where flnep0 is the friction mean effective pressure and imepPi0 the indicated

mean effective pressure of the motored engine. As long as no high load oper¬

ating points are considered, the measurements and simulation results of the

conventional engine bmep should follow the same line. In the case of the sec¬

ondary valve equipped engine some differences are expected. At a first glance

one may expect a parallel shift of the Willan's line towards higher bmep but

due to the deterioration of the burn rate at low fuel mep the fuel benefits of

secondary valves are reduced in that region. Even though at zero fuel mep the

pumping loop mean effective pressure of a secondary valve equipped engine is

higher than for a conventional engine, the slope of the curve in the secondaryvalve case remains smaller than the slope of the Willan's line for the conven¬

tional engine since the burning rate of an engine with early intake valve clos¬

ing scheme is worse than the one of a conventionally throttled engine.

bmep^ sv> bmepQf mR (5.2)

eSV~eTHR (5.3)

5.1 Analysis with Willan's Approach 59

Figure 5-1: Relevant part of the fuel map for the MVEG-95 driving cycle.

At peak power no fuel consumption difference is expected since the two load

controls become identical. Thus the trace bends down towards the maximum

bmep of the conventional throttled engine. A scheme of the expected curve of

bmep in function of fuel mep of a secondary valve equipped engine and the

Willan's line of a conventional engine are shown in Figure 5-2. In conclusion,

the linear Willan's approach should not be used for fuel consumption analysisof an engine equipped with secondary valves because linear correlation

between bmep andfuel mep is not guaranteed a priori. Such an analysis needs

to be seen as a first order approach. Nevertheless the Willan's method remains

a convenient way of cross checking the simulation results and the equivalent

measurements of a conventional engine.

The sub-plots of Figure 5-3 and Figure 5-4 show Willan's lines of the exper¬

iments at different engine speeds. It is seen that the measurements of the con¬

ventional engine and the corresponding simulation results show the expected

aligned pattern. The agreement between measurement (triangles) and simula-


bmep

— Production Engine- - Base Engine with Secondary Valves

bmepo

fuel mep

Figure 5-2: Schematic Willan's lines for the two different approaches.

tion (squares) bmep of the conventional throttled engine were within an

acceptable error range of ±5% bmep. These inaccuracies are caused by sim¬

plified modeling of the engine process and by damping pressure waves. A

lumped parameter engine model acts as a low pass filter and can therefore not

completely resolve high intake and exhaust pressure oscillation frequencies.

Since the exact algebraic form of the curves which should be fitted throughthe results of the secondary valve equipped engine is not known, no traces

were drawn in the sub-plots of Figure 5-3 and Figure 5-4. Nevertheless, the

bmep's of the secondary valve equipped engine lie above the Willan's lines of

the measurements and simulations at almost every engine speed and load.

Therefore it can be expected that secondary valves reduce the fuel consump¬

tion when deployed in an early valve closing mode. At operating points in the

vicinity of zero bmep, the burn delay deterioration effect is clearly visible

throughout the whole speed range. It can best be seen in the lower right sub¬

plot of Figure 5-3 at 2000 rpm: the burn rate delay at zero bmep reaches as

much as 14 degrees crank angle, reducing completely the benefits of the sec¬

ondary valve. Only at higher loads the burn delay is lowered, improving bmep.

5.2 Fuel Map 61

x 1 q5 n = 800 [rpm]

CO o

n ^

-o 0

0.5 1.5

x10

x 1Q5 n = 1500 [rpm]

CO o

-ft 0

-2

A

30j&

<tâ

0.5 1 1.5

/we/ mep [Pa] x 10

105 n = 1000 [rpm]

0

-20.5 1.5

105 n = 2000 [rpm]

x10

A Measured

<> SV

a THR

0.5 1 1.5

fuel mep [Pa] x 10

Figure 5-3: Willan's lines for different speed (THR and SV from simulation).

5.2 Fuel Map

The fuel map of the engine with secondary valves is directly drawn with the

results obtained by the simulation. By superimposing the conventional throt¬

tled engine fuel map on top of the secondary valve equipped engine map

(Figure 5-5) the fuel consumption benefits become visible. The conclusion is

that at a specific engine speed secondary valves allow higher torques for a

given fuel consumption (or at a given torque the fuel consumptions is

reduced).


1Q5 n = 2500 [rpm]

CO o

CL d

5T

~ö o/

-0"p

0.5 1.5

x i Q5 n = 3500 [rpm]

x10

CO o

^ 0

...: .....a

: OX :

OJsT

0^

0.5 1 1.5e

fuel mep [Pa] x 10

x1Q5 n = 3000 [rpm]

0

-2

J&

4P

'/*

<$*

0.5 1.5

x10

x i q5 n = 4000 rpm

fuel mep [Pa] x 10

Figure 5-4: Willan's lines for different speed (THR and SV from simulation).

The computed relative fuel consumption reduction map of Figure 5-6 sum¬

marizes the result of all the thermodynamic engine cycle simulations. It was

obtained by smoothing the fuel consumption results of the simulation showen

in Figure 5-5 and computing the relative fuel savings Am, at every operatingpoint.

Amf =

m.r -mr +lf, sv f thr

100m

(5.4)7, thr

Where mrsv

is the fuel consumption of the secondary valve equipped engineand mr

fnrthat of the conventionally throttled engine.

5.2 Fuel Map 63

500 1500 2500 3500 4500

Engine Speed [rpm]

Figure 5-5: Fuel map of the conventional and secondary valve engine [kg/h].

500 1500 2500 3500 4500

Engine Speed [rpm]

Figure 5-6: Relative fuel consumption change in percent.


The conchiform shape of the iso-fuel-consumption-change curves of

Figure 5-6 shows clearly the predicted effect, i.e. that the burn rate deteriora¬

tion at low load reduces the benefits of early valve closing. The deterioration

is more significant at lower engine speed than at higher ones because of the

diminution of the in-cylinder turbulences during the intake and the compres¬

sion stroke when the piston speed is low. Best fuel consumption benefits are

located at approximately 30 Nm over the entire measured engine speed range.

The reduction of the fuel savings at higher torque is due to the fact that less

pumping work can be saved.

5.3 Residual Gas Fraction Map

The simulation results show a reduction of residual gas fraction at all ana¬

lyzed operating points. By the nature of the process, residual gas fraction dif¬

ferences between the conventional throttled engine and secondary valve

equipped engine can vary substantially from one operating point to the other.

They mainly depend on the intake and exhaust pressure traces used as inputsfor the simulations. Nevertheless, the residual gas fraction map shows a dis¬

tinctive trend: the lower the engine speed and torque is, the higher is the resid¬

ual gas fraction difference. The residual gas fraction differences depend on

load which shows a direct relation to the manifold pressure. With secondaryvalve load control, manifold pressure stays almost atmospheric. In contrast,

when the load is controlled with a throttle plate the manifold pressure can be

as low as 0.3 TO Pa. The amount of burned gas flowing back into the intake

runner is thus higher, as discussed in Chapter 3. With increased engine speed,the gas flow in the intake and exhaust system develops important inertia

effects, helping scavenging the cylinder. This reduces the residual gas fraction

and hence the benefits of a secondary valve on residual gas fraction diminsh.

The residual gas fraction reduction is given by

sv thrIC- C\

x = — (5.5)

xthr

5.4 Burn Delay Map 65

500 1500 2500 3500 4500

Engine Speed [rpm]

Figure 5-7: Relative change of residual gas fraction in percent.

5.4 Burn Delay Map

The burn delay is strongly dependent on the reference pressure and tempera¬

ture in the cylinder during the compression stroke as shown in Chapter 4.3.4.

These two values shift the 50% conversion point of the burn rate towards BDC

if they are lower than the reference values. This phenomena can be seen in

Figure 5-8.

The burn delay map is given by

XB(50%)=

XB(50%), thr~ XB(50%), sv^^

Some correlation on engine speed exists, mainly determined by the effects of

residual gas rate change. The Csallner approach used in this work does not

include engine speed dependencies because the conventional engine and the


secondary valve equipped engine are simulated at same engine speeds. The

only engine parameter that reduces the burn rate delay is the diminution of the

residual gas fraction, as calculated and given in Figure 5-7. With higher speedthe residual gas fraction decreases compared to the conventional throttled

engine. Therefore, the burn rate delay increases at higher engine speed. The

drop at approximately 2800 rpm throughout all loads is due to ram effects.

All in all, the tendencies are as expected and show the trend of how second¬

ary valves alter the burn rate delay when used in an early intake valve closure

mode.

5.5 Secondary Valve Timing Map

Finally, the secondary valve timing map is shown in Figure 5-9. This figureleads to two conclusions expected from previous discussions. First, the closingtime of the secondary valve moves towards BDC with increasing load. Sec¬

ond, a closing time dependency on engine speed at same load because second¬

ary valve closing transitions are related to time and not to crankshaft degrees.At the highest analyzed loads a closing time at approximately 430 degreescrank angle is required. The parasitic volume is responsible for the fact that

the closing time is not delayed to higher crank angle degrees. At low loads the

closing time can be less than TDC, i.e. 360 degrees crank angle. Once againthe parasitic volume makes this possible: without parasitic volume the closingtime would not move below TDC because no fuel-air mixture enters into the

engine.

As mentioned, the closing duration depends also on engine speed because

relative to the induction stroke duration the closing transition duration of the

secondary valve increases. This implies more flow friction, reducing the effi¬

ciency of the secondary valve. To compensate for this effect the secondaryvalve is closed earlier with higher engine speed at a given load.

5.5 Secondary Valve Timing Map 67

500 1500 2500 3500 4500

Engine Speed [rpm]

Figure 5-8: Burn delay map [°CA].

500 1500 2500 3500 4500

Engine Speed [rpm]

Figure 5-9: Closing time of the secondary valve [°CA].

Blank leaf

69

Chapter 6

Model Sensitivity Analysis

Three model parameter variations were analyzed: varying the parasitic vol¬

ume, the intake air temperature and the secondary valve transition duration.

The aim of the first two sensitivity analysis is to show how secondary valve

closing times and fuel consumption are affected. The last variation investi¬

gates two different secondary valve operation modes, allowing to run the

engine at low loads, almost independently of the parasitic volume size.

All simulation results shown in this chapter are based on the model intro¬

duced in Chapter 4 and at an engine operating point of 2000 rpm and

3.5-10 Pa imep.

6.1 Variation of the Parasitic Volume

As shown in Chapter 3, the closing time is strongly depended on the para¬

sitic volume. A larger parasitic volume shifts the closing time of the secondaryvalve towards TDC or even earlier as in the case of very low loads and consid¬

erable parasitic volume. If the closing time of the secondary valve is earlier

than a certain value, the engine load remains constant and can not be reduced

any further. The reason is that the parasitic volume is filled with fuel-air mix¬

ture which will be partially inducted during the next stroke when this gas is

expanded during the intake stroke. Five in-cylinder pressure traces versus in-

cylinder volume are drawn in Figure 6-1 - all at 2000 rpm and 3.5-10 Pa

1. See Chapter 3 for further details

Chapter 6 Model Sensitivity Analysis

Figure 6-1: In-cylinder pressure at several secondary valve closing times.

imep but with different parasitic volumes. The conclusions made in Chapter 3

are verified: increasing the parasitic volume moves the closing time of the sec¬

ondary valve to earlier crank angle degrees. Furthermore, the in-cylinder pres¬sure slope between secondary valve closing and BDC is reduced

proportionally to the increase of parasitic volume. This is due to the polytropicprocess: in-cylinder pressure relates to the cylinder volume plus the parasiticvolume.

p-V1 = p.(Vcy+Vpf = const. (6.1)

Where n is the polytropic exponent.

The closing time as a function of parasitic volume is shown in Figure 6-2.

The x-axis is labeled also with the parasitic volume related to the clearance

volume Vc. Without claiming a linear relationship between secondary valve

closing time and parasitic volume, a straight line was fitted through the results

6.1 Variation of the Parasitic Volume 71

410

400 -

< 390Ü

CD380

C

'co

_oO

>CO

s 370

360

350

340

1

Ix1

^

1.1

Ü>

>\D

- - -" " -

"1

1

1Q_""""""

11

0 1.64

V [m3]p

L J

3.28

V / VP c

4.91

4 5-4

x10

6.55 8.19

Figure 6-2: Secondary valve closing time as a function of parasitic volume.

of the simulation to point out the trend. If the parasitic volume is set to zero,

the secondary valve system does not differ from a variable intake valve system

(VIVC). Therefore, values given at a zero parasitic volume are equal to those

obtained with directly actuated intake valves used for early intake valve clos¬

ing schemes.

The analysis of the fuel consumption leads to the following conclusion:

increasing the parasitic volume, increases the pumping work as well. Conse¬

quently, the pumping loop benefits are diminished. On the other hand, in-cyl¬inder pressure and temperature decrease are less severe leading to improvedburn rates. The former effect is more relevant and raises the fuel consumptionas much as 3.7% compared to the zero parasitic volume case.

72 Chapter 6 Model Sensitivity Analysis

9.33

9.26

S 9.20

9.13

x10

9.07

c

.o

Q.

E

COc

o

8.93 -

8.87

I

..

..I......

. . .

.J^.. O-

ID yf

I

Ü>

>

I

\""/£

I

0

V [m3]P

Jx 10

-4

Figure 6-3: Fuel consumption as a function of parasitic volume.

A standard engine head equipped with secondary valves is expected to have

a parasitic volume of approximately 2.5 times the clearance volume. The sec¬

ondary valve device cannot reduce this volume, otherwise peak engine effi¬

ciency is reduced due to deterioration of volumetric efficiency. Assuming a

parasitic volume of 2.5 times the clearance volume, a fuel consumptionincrease of 1.1 % compared to a system without parasitic volume is expected,see Figure 6-3.

1. corresponding to the volume of the intake runner from intake port to

intake valve

6.2 Variation of the Manifold Temperature 73

6.2 Variation of the Manifold Temperature

Another way of reducing pumping work is to increase the air temperature in

the manifold and intake runner. In the case of a throttled engine this corre¬

sponds to a higher manifold pressure for the same imep. A secondary valve

equipped engine, however, maintains the manifold pressure at a value slightlyless than atmospheric, but with higher fuel-air temperature the closing time of

the secondary valve needs to be set later to achieve the same imep. This depen¬

dency is demonstrated in Figure 6-4 at an engine operating point of 2000 rpm

and 3.5-10 Pa imep.

The fuel consumption is only slightly affected throughout the range of fuel-

air temperature analyzed. Moreover the calculated values lie within the range

of the tolerated simulation error. The reason is that the closing time of the sec¬

ondary valve varies only by about 10 degrees crank angle over the analyzed

temperature range. A slight improvement in fuel consumption at higher tem¬

perature due to the better burning of the cylinder charge is expected but could

not be demonstrated, Figure 6-5. This can have several causes. For example,the simulation did not optimize the ignition timing for MBT even thoughshorter burning duration calls for later ignition timing. Another possible rea¬

son is the increased compression work at higher in-cylinder fuel-air mixture

temperature because of the temperature depending specific heat at constant

volume. At a certain temperature level the pumping work reduction and fuel-

air mixture pre-heating is balanced with the increased compression work. Fur¬

thermore heat transfer from the cylinder wall to the in-cylinder fuel-air mix¬

ture can also reduce the benefit of pre-heating fuel-air mixture since the

temperature difference between gas mixture and wall is reduced. In some

cases the heat transfer might even be inverted during the early phase of the

engine cycle.

In conclusion, pre-heating the fuel-air mixture seems not to significantly

improve fuel benefits obtained with an engine equipped with secondaryvalves. Moreover, heating up the fuel-air mixture needs a bulky heating sys¬

tem which at higher engine speed will not be able to heat the fuel-air mixture

fast enough. Finally, problems with transient control are likely to arise.


300 450350 400

Manifold Temperature [°K]

Figure 6-4: Secondary valve closing time as a func. of manifold temperature.,-4

9.1x10

-Ï2

^9.05

o"4—"Q.

E

CO

c

o

V9

CD^

LL

8.95

I

TDC = 360° CA

i

p

I

d

Su

'

p y

p

p

... ., S-u

-

-

yS D;

•

;

.

300 350 400

Manifold Temperature [°K]450

Figure 6-5: Fuel consumption as a function of manifold temperature.

6.3 Variation ofSecondary Valve Closing Times and Durations 75

6.3 Variation of Secondary Valve Closing Times and

Durations

Two secondary valve modes were considered: one uses a secondary valve

with fast transition duration whereas the other uses the performance of the one

used during the experiments. As demonstrated in Chapter 3, the parasitic vol¬

ume might prevent the engine from running at very low load if the secondaryvalve is used in the conventional way - usually the secondary valve opens dur¬

ing the high pressure loop and is closed somewhere during the intake stroke.

With this method, the fuel-air mixture flows into the parasitic volume as soon

as the secondary valve opens again. Notice that after the intake valve closes

the pressure in the parasitic volume remains much lower than atmospheric.The fuel-air mixture in the parasitic volume will expand into the cylinder after

the secondary valve closes, producing engine torque even if the secondaryvalve was shut before induction stroke start. This conventional mode of opera¬

tion (case a)) was used as base and is shown in Figure 6-6. The other two

a) Base

b)

c)

1

0.5 h

0a

CDco

-§ -200

>en 1

°0.5c

8- °

> -200

1

0.5

0

2.5 ms

0 200 400 600

I I I

"

waiting period "

V

—*

ö.^ö ms

i i i

0 200 400 600

1.25 ms'

-200 0 200 400

Crank Angle [deg] (0,360 = TDC; 180,540 = BDC)

600

Figure 6-6: Secondary valve operation modes.


cases shown in the same figure represent the other mode of operation investi¬

gated in this section: the secondary valve remains closed during the high pres¬

sure loop (case b)) and does not open again as done in the base case. The

crank angle at which the secondary valve starts to open is found by calculatingbackwards from the required secondary valve closing time. It is assumed that

the secondary valve actuator requires half of the transition duration as waiting

period before being ready to close the secondary valve again. Therefore a total

cycle needs 6.25 ms: 2.5 ms to open, 1.25 ms for the waiting period and

another 2.5 ms to close the secondary valve. The device used for the experi¬ments did not allow such fast cycle duration since from the triggering signal to

the motion start of the secondary valve a 6 ms delay was inevitable. In the

third mode (case c)), the same was done but with a (not yet realizable) transi¬

tion duration of only 0.5 ms, leading to a cycle duration of 1.25 ms.

The idea of this particular secondary valve operation mode is to let residual

gas flow back into the intake runner and thus into the parasitic volume after

the intake valve opened. Therefore zero load or less can be achieved, if the

secondary valve closes exactly when the parasitic volume contains only resid¬

ual gas. With this technique, no fuel-air mixture flows into the cylinder. Since

perfect mixture of residual gas and fuel-air mixture was implemented into the

engine model, some fuel-air mixture mass always remains in the parasitic vol¬

ume during the backflow process. Consequently the fraction of burned fuel-air

mixture mass and total gas mass in the parasitic volume never reaches one.

Nevertheless it is still significantly increased compared to the conventional

case a).

Figure 6-7 and Figure 6-8 show the results when the secondary valve closingtime is varied from 250 to 600 degrees crank angle and imep is computed with

the engine model of Chapter 4. The valve trajectories used are those of

Figure 6-6.

In case b), it was seen that the overall open duration of the secondary valve

was to long to achieve zero load. Figure 6-7 shows this effect: starting at

approximately 360 degrees crank angle the load does not decrease anymore.

1. case b) and c)

6.3 Variation ofSecondary Valve Closing Times and Durations 11

Only when the secondary valve closing time is set to late, load starts to drop.This is due to the intake valve closing and the increased pressure after the pis¬ton starts to move up again, stopping fuel-air mixture to flow into the cylinder.

Case c) however, with a very fast secondary valve is capable of reducing the

load as anticipated. The open duration of the secondary valve is short enoughto allow only a small amount of fuel-air mixture to enter the parasitic volume

and hence the cylinder. Likewise the decrease of load when the secondaryvalve closing time is set to late is due to the same effects as mentioned above,

case b).

Since this method of reducing load depends on the residual gas fraction of

the gas in the parasitic volume, an increase of residual gas should be observed

with a peak at lowest imep. This can be seen in Figure 6-8. In case c), the

residual gas fraction peak coincides with the decrease of imep. On the con¬

trary, in case b) it is clearly visible that the residual gas fraction is not

increased and therefore load not reduced.


12x

'

101 i 1 ! ' !

TDC:= 360° CAy

10

8

6

'ccÉL 4

1 2

- l

/

y/

////

\

\

\

\

. A"

\ \

\ \

r

y

-/ /

0 -

-2 - case a)- - case b)

-4 - case c)i i . i

. i

250 600300 350 400 450 500 550

Secondary Valve Closing Time [°CA]

Figure 6-7: Mean effective pressure - different secondary valve modes.

0.7

TDC = 360° CA

- case a)- case b)— case c)

250 300 350 400 450 500 550

Secondary Valve Closing Time [°CA]

600

Figure 6-8: Residual gas fraction in the parasitic volume.

79

Chapter 7

Experiments

To analyze the impact of secondary valves on engine performance a proto¬

type device carrying four actuators and secondary valves was realized and

mounted between the inlet manifold and the engine. In this way both, back-

flow and load control experiments could be performed. The operating range of

the pneumatic actuator was restricted by reliability and performance limita¬

tions. A part load fuel map was measured with the same engine without sec¬

ondary valves. This covered the relevant operating points of an engine used

during the MVEG-95 driving cycle (80 operating points were measured).

Finally, driving cycle experiments with a vehicle emulation system deter¬

mined the fuel consumption as a function of time. All the results were used for

the validation of the simulations or as starting points for estimations beyondthe range of the feasible measurements.

7.1 Static Measurements - Engine with SecondaryValves

Measurements of an engine equipped with secondary valves and of a con¬

ventional engine provided the necessary information for the experimental

analysis of the impact a secondary valve system has on the engine perfor¬mance. The secondary valve was tested at two different engine speeds and

several loads.

80 Chapter 7 Experiments

7.1.1 Measurement Setup

The secondary valve experiments were performed on a Ford Zetec DOHC 4-

cylinder 2.0-liter engine coupled to a static dynamometer. The dynamometer

(Zoellner/AVL brake) was equipped with a standard measurement system, with

torque, speed, oil temperature and pressure sensors. All auxiliary engine

devices, as found in the series production car, were connected to the engine,but some, e.g. alternator and power steering pump, were not loaded. The ther¬

mal management of the engine used the oil and water pump of the conven¬

tional engine. Finally, the engine was mounted on a test rig and coupled to the

brake without a gearbox. For the roughness control experiments the enginewas supplied with specially manufactured camshafts from Dunnell Engines.These camshafts had particularly large valve overlap timings leading to

increased peak power at higher speed regimes with inacceptable roughness at

low part load compared to standard camshafts. All load control experimentswere performed using the standard Ford camshafts. Fuel injection and ignitionwere controlled using a custom electronic engine controller developed byDunnell Engines. The air-fuel ratio was set manually to stoichiometric condi¬

tions during all experiments and ignition adjusted to obtain maximum break

torque. The secondary valves were controlled with a custom electronic con¬

troller based on an Altera FPGA, which was configured through a serial com¬

munication link by the PC. Figure 7-1 contains more details on the set-up. All

important slow variables (less than 10 Hz bandwidth) such as intake tempera¬

ture or atmospheric pressure, were recorded using a PC, based on Labview

data acquisition hard- and software from National Instruments. High band¬

width sensors as required for accurate pressure measurements were linked to a

camshaft triggered data logger (KRENZ, 12-bit quantization and 0.1 degreescrankshaft sampling-rate) capable of logging each channel 3600 times per rev¬

olution. Fast response Kistler pressure sensors were used to measure the

intake and in-cylinder pressures.

7.1.2 Secondary Valve Prototype

An aluminium interface block was designed and built in two halves to carry

the four secondary valves and their actuators. This block fitted between the

cylinder head and the original intake manifold of the engine. Injectors and

throttle are fastened on the manifold. Figure 7-3 shows a cross-section of this


Data Logger max. 10 Hz

Brake 64

toData Logger3600/rev

1 Speed set point

2 Torque

3 Exhaust pressure

4 Exhaust temperature

5 Trigger signal

6 Exhaust temperature

7 Oil temperature

8 Oil pressure

9 Intake pressure

10 Intake temperature

11 Throttle position

12 Throttle position set point

13 Ambient pressure

14 Ambient temperature

15 Air flow

16-19 Secondary valve signal

20 Ignition

21 Crank sensor one

22 Crank sensor two

23 Ignition

24 Intake pressure

25 In-cylinder pressure

26 Injection

Figure 7-1: Measurement setup of the test bench.


Figure 7-2: Engine with secondary valve device.

block, including a secondary valve. A general view of the block, showing how

the halves fit together, is given in Figure 7-4. In this view the right hand side

attaches to the engine cylinder head and the conventional manifold is bolted

onto the left hand side.

The actuators are partially visible on the lower right hand side of Figure 7-4.

Figure 7-5 shows the four actuators inserted into the machined aluminium

block, the passages of the individual intake runners from the manifold and the

slots for the injectors. Figure 7-6 shows underneath a picture from the mani¬

fold-side half of the block, in which the four valves can be seen, and the upper

picture shows the engine-side half with the valve seats visible at the bottom of

the machined recess. The actuators were commercially available electro-pneu¬

matic linear actuators from Eugen Seitz A.G. (Figure 7-7). which were oper-


210 mm

Figure 7-3: Secondary valve carrier block - technical drawing.

ated with higher than rated pressure and control currents in order to increase

their speed. The camshaft position pick-up sensor conflicted with the carrier

block and had to be slightly shifted: no other alterations were made to the

engine. The main objectives of the design were to minimize both, intake flow

resistance and parasitic volume between the secondary valve and the conven¬

tional intake valve. A pressure sensor was fitted to measure the pressure in the

trapped volume during the experiments. The moving elements of the valve

(Figure 7-8) were machined from a high-strength lightweight plastic (Polya¬mide 4.6, ISO 1043). The dynamic behavior of the secondary valves was mea¬

sured, and the system was found to have a 3.5 ms latency time, compensated


Figure 7-4: Secondary valve carrier block - full view.

for by pre-triggering, and an opening time between 2.5 and 9 ms - dependingon whether the valve is being opened or closed, and on the pressure difference

across the secondary valve. The weight of one secondary valve was approxi¬

mately 25 g. The system worked well and fulfilled the design specifications.No claim is made that this specific design can evolve into a production sys¬

tem; the durability, cost and power of this approach do not meet the require¬ments of such systems. However, the system meets the requirements for a first

experimental investigation, as needed in this work.

No effort has been made to reduce the energy consumption of the secondaryvalve actuator which uses approximately 8 J per cycle and cylinder. This

includes electrical and pneumatic energy for the activation of the valve for one

complete cycle, i.e. opening and closing.

7.1 Static Measurements - Engine with Secondary Valves

j&: ; li

Figure 7-5: Secondary valve carrier block - manifold side.

Figure 7-6: Secondary valve carrier block - flow channels.


Figure 7-7: Pneumatic actuator.

20 mm

Figure 7-8: Secondary valve.


7.1.3 Roughness Improvements

There are many different methods of characterizing an engine's roughness,for example based on measuring the variability of peak cylinder pressure or of

engine rotational speed. One commonly used procedure is to determine the

standard deviation of the net energy generated during the compression and

expansion strokes of at least one cylinder. This is equivalent to computing the

variability in imep from cycle to cycle but without including the pumping loopwork part. This is the most basic measure of an engines roughness due to mis¬

fire and poor combustion; although roughness due to cylinder-to-cylinder vari¬

ability is neglected. Mathematically the measure of engine roughness can be

expressed as a standard deviation

STD = ii = i

\2n

(7.1)n-1

with h denoting the high-pressure part of the cycle from BDC to BDC.

Speed and load combinations were chosen without activating the secondaryvalve but by using the throttle. The secondary valves were then set to open

early enough when first switched on, thus no back-flow control occurs. After¬

wards the opening time of the secondary valves was incrementally delayed,and as backflow control became effective the engine torque started to increase.

The fueling had to be adjusted every time the timing of the secondary valves

was modified by changing the injection period until the air-fuel ratio was sto¬

ichiometric. Again the best secondary valve timing was considered to be justbefore the torque began to decrease. This decrease begins when the secondaryvalves remains closed during the start of the intake stroke, resulting in load

control by late intake valve opening. When the best timing was reached, oper¬

ating conditions were recorded. This data was compared to data recorded

without operating the secondary valves but with the same load and speed.

Figure 7-9 illustrates the operation of the secondary valve in roughness con¬

trol. It shows the pressure upstream of the main intake valve, downstream of

the secondary valve respectively, without (solid) and with (dashed) operation


x 10

12 -

£10CD

CO o

CO o

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® 6c

DC

cd 4

CO

0

Without backflow control

With backflow control

* SVOpenh

,' \ ,A SV Closed

SV Opening Signal SV Closing Signal

./

TDC BDC TDC BDC

200 400 1000600 800

Crank Angle [deg]

Figure 7-9: Roughness experiments (1000 rpm; 65 Nm) - intake pressure.,4

15x10

OB

CD10

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- - With backflow control: i :0

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Cylinder Volume [m3] x 10-4Figure 7-10: Roughness experiments (1000 rpm; 65 Nm) - pumping loop.

7.1 Statte Measurements - Engine with Secondary Valves 89

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CD-50 -

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_c

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+ Without backflow control1 1 1 1 i i i

0 10 15 20 25

Cycle Number

30 35

Figure 7-11: Roughness experiments: standard deviation of combustion loop.

of the secondary valves. The loads are equal in both cases. The first arrow

indicates the electronic triggering to open the secondary valve; 6 ms later,

marked with a filled gray box, the secondary valve is fully open. The pressure

in the parasitic volume rises to the exhaust pressure until the secondary valve

opens, after that the in-cylinder pressure rapidly falls to manifold pressure. It

is seen that without the secondary valves the manifold pressure needs to be set

higher to obtain the same load to compensate for charge dilution by exhaust

gas. Figure 7-10 shows the pumping loop portion of the pressure-volume dia¬

gram. It can be seen that the manifold pressure is higher in the case without

back-flow control, as stated before.

The in-cylinder pressure measurements showed that better filling due to

backflow prevention should decrease the cycle-to-cycle variations in torquedue to bad burning. This can be shown by plotting the standard deviation of

the indicated work of the high pressure part of the engine cycle for a consecu¬

tive sequence of cycles as done in Figure 7-11. In a typical operating point the

standard deviation of the high pressure work without backflow control is 98 J


whereas with backflow control it can be reduced to 12 J,

(Equation 7.1 on page 87). The delivered power of the engine is the same in

both cases but in the case with secondary valves no misfiring occurs and cycleto cycle pressure variation is less because of better charge filling. In contrast,

without secondary valves misfires happen often and the pressure variability is

substantially increased. The roughness of the engine at different loads, with

and without backflow control and with aggressive or standard camshafts at

980 rpm is shown in Figure 7-12. At higher loads, as expected, dynamiceffects purge the cylinder resulting in less roughness. By using the secondaryvalve major roughness improvements were obtained at all engine speeds com¬

pared with the aggressive camshaft. At higher loads even a 50% improvementover the normal Ford camshafts and a four fold improvement over the aggres¬

sive camshafts could be demonstrated. At low load the reduction of roughness

is 2.5 times, achieving a level approximately comparable with the standard

Ford camshafts. Figure 7-13 shows the measurements at 1600 rpm. Again the

roughness is substantially reduced by the action of the secondary valve, in this

case always to a level as low as or lower than that obtained with the standard

camshafts.

7.1.4 Load Control

Throttleless load control was implemented using the secondary valve in an

early intake valve closure mode as described in the previous chapters. The sec¬

ondary valve was rapidly shut during the intake stroke while the piston was

still moving down the cylinder, thus limiting the air and fuel charge inducted.

In this mode of operation the engine power is controlled by varying the clos¬

ing timing of the secondary intake valve. The secondary valve was able to

control the load as anticipated. Shown in Figure 7-14 is the pumping loop por¬

tion of the pressure-volume traces recorded with conventional throttled load

control and with throttleless secondary valve load control. The load in each

case is virtually the same - about 30 Nm. In the throttled case the pumping

loop shows the conventional trajectory: in particular, when the intake valve

opens the pressure falls abruptly to intake pressure, approximately0.4-10 Pa, where it stays fairly constant during the induction stroke. In the

throttleless load control case the pressure at the start of the intake stroke is vir¬

tually atmospheric pressure - there is no throttle, so the manifold pressure is

atmospheric less a slight drop due to losses in the intake system. When the


70

60

50

$40CDc

"130o

rr

20

10 -

0

i i i

---

A Aggressive Camshafts, Backflow Control0 Aggressive Camshafts, No Backflow Control

Standard Camshafts

KV

<i'...5i....

.... - - -Ak

. ....A. ^S-« 0

nn d ^""""-

n ra—, \^<>°

D U""**&"

20 40 60

Torque [Nm]

80 100

Figure 7-12: Roughness experiments at 980 rpm.

50 -

40 -

«30CDCSZ

a>

I 20

10

0

i I I I I

A Aggressive Camshafts, Backflow Control<> Aggressive Camshafts, No Backflow Control

Standard Camshafts

0

0\ <^\^

^^0^0

/\Ci

'

fAJUTA n A a z\

A; Ä Ä

20 30 40 50

Torque [Nm]

60 70

Figure 7-13: Roughness experiments at 1600 rpm.


15x10

10-

CÖ

cl

CD

CO

COCD

CDT3

c

% 5oc

0

\\\\\\\\

-,_,^ \ \

-^

"""

^ -~

i

\

\

\\\\\\

y\[

!

Throttle load control- - SV load control

i i

0

Volume [rrr] x10

6

-4

Figure 7-14: Load control experiments: pressure loop (980 rpm; 30 Nm).

secondary valve closes, in this example at a cylinder volume of approximately15-10 m

,the pressure in the cylinder falls polytropically as the volume

increases until the end of the intake stroke, when it is comparable to that in the

throttled case. At this point the main intake valve closes and the pressure rises

as in the standard case during the intake stroke. The pressure rise is steeperthan the fall during the induction stroke because of the presence of the para¬sitic volume between the secondary valve and intake valve. The area between

the two induction stroke lines is equal to the pumping loop reduction - up to

70% of the standard pumping work. The behavior of the process is further

illustrated in Figure 7-15, where the runner pressure in the region of the para¬sitic volume is shown during the intake event. In the conventional throttled

case this is essentially a constant pressure trace at approximately 0.4-10 Pa,

apart from small fluctuations due to intake pressure wave dynamics. In the

throttleless case, the pressure is near atmospheric until the secondary valve

closes. The runner pressure then follows the in-cylinder pressure, decreasingalmost isentropically and raising slightly just before the intake valve closes


15x10

CO

CL

CD

5 10COCO

CD

CDcc

cr

CD

CO

SV Closing Signal

Without load control

With load control

SV Closed

SV Opening Signal

SV Open

TDC BDC TDC BDC

0200 400 600 800

Crank Angle [deg]

1000

Figure 7-15: Load control experiments: intake pressure (980 rpm; 30 Nm).

(after BDC). After the intake valve has closed the fuel-air mixture between the

secondary valve and the main valve is trapped and remains at an approxi¬

mately constant pressure. When the secondary valve opens, to be ready for the

next stroke, manifold pressure returns to ambient value. The gradual rise is

due to heat flow phenomena. Note that at the start of the compression stroke

the in-cylinder conditions are not identical in the two cases, as discussed in

Chapter 4.

For example at a low operating point pumping work saving of 70% can be

realized but the fuel economy improvement is only in the order of 4%. The

important difference between these two values is the conflicting effects of

throttleless load control: on the one hand the pumping work is reduced, on the

other hand, the burn rate slowed down because the charge is cooler and the in-

cylinder turbulence is reduced.


6.5x10

co 5.5

.1 5

4—•

Q.

| 4.5

COc

8 4

CD

1^3.5

2.5

l

0 THR

p SV.

iff

1 1

0.5 1.5 2 2.5

bmep [Pa]

3.5

x10^

Figure 7-16: Measured fuel consumptions at 980 rpm.

Figure 7-16 shows the fuel consumption observed during the experiments

using conventional throttling and throttleless load control as a function of load

at 980 rpm (solid: conventional throttle). It can be seen that in the case of

throttleless load control fuel consumption is reduced (squares), although with

higher load the difference diminishes. This was expected because at full load

the throttle is completely open and so no pumping work can be saved.

7.2 Static Measurements - Conventional Engine

For the static conventional engine measurements another test bench was

used. The mode of operation was maintained as in the former experiments and

since the main engine setup was the same as for the secondary valve experi¬ments the result are comparable.

7.3 Dynamic Measurements - Conventional Engine 95

7.2.1 Measurement Setup

The major difference from the first experiments is the test bench which was

now a dynamic test bench of APICOM, type SM 200 L-4. The speed was con¬

trolled by the test bench and the load with the throttle. No secondary valves

were mounted to the engine and instead of using the electronic control unit

from Dunnell Engines the original Ford EEC was used. For precise pressure

measurements a water cooled Kistler pressure sensor was mounted into cylin¬der number two. Since dynamic intake and exhaust pressure measurements

were needed for the engine simulation an additional pressure sensor was

inserted shortly after the exhaust port. Fuel measurements were made with a

flow meter from Quickly AG (device name Flowtronic). The entire measure¬

ment setup was automated to run without manual adjustments, the measure¬

ments triggered as soon the engine was running at steady-state and after

completion of the measurement the next operating point was automaticallyset. To allow zero break mean effective pressure at every speed it was neces¬

sary to block the idle-speed bypass of the throttle.

7.2.2 Result of the Static Measurements

Next to the pressure measurements as described in Chapter 4 and Chapter 5

the fuel map obtained is shown in Figure 7-17. This map represents only the

lower left part of the entire engine operating range as shown in Figure 5-1.

This range is sufficient for the driving cycle simulation because higher loads

and speeds are not needed in the MVEG-95.

7.3 Dynamic Measurements - Conventional Engine

The concept of the driving cycle emulation is explained in detail in [41]. The

vehicle emulation on the dynamic test bench was implemented using a real¬

time controller from dSpace. The code was written in Matlab/Simulink and

tested on a conventional PC. All relevant parameters such as inertia of engine

components, etc. were identified with preliminary tests. The code emulates the

vehicle, the transmission and the driver in real-time with a maximum sam¬

pling time of one millisecond. The correct throttle position is computed from

the actual torque, speed and time and the torque set point required to power


500 1500 2500

Engine Speed [rpm]

3500 4500

Figure 7-17: Measured fuel consumption map o the production engine [kg/h].

the vehicle at this moment. It is important to notice that the car driver is simu¬

lated with a set of different controllers activated one at a time by a speciallogic. This logic decides which of the controller is active depending on the

task the driver needs to perform - for example gear change, braking or acceler¬

ation.

The emulated car is a Ford Mondeo with a manual five speed gear-box. All

important values for the driving cycle simulation can be found in the appen¬

dix. The driving cycle used is a MVEG-95 cycle as prescribed by the Euro¬

pean norm [49].

To allow comparison with the quasi-static driving cycle simulation described

in Chapter 8 no cold starts were measured. The engine was first switched on

and cycles were only measured after the oil temperature remained constant.

The result are given and discussed in detail in Chapter 8.

97

Chapter 8

a case study: mveg-95

Fuel Consumption

Driving a test cycle under real conditions is a delicate task and gives differ¬

ent results even with the same car. Influences are weather, road condition and

most importantly the driver himself [45] [46]. Therefore the human driver is

often replaced by a robot and the vehicle fixed on a chassis dynamometerrather than driven on the road. The next step for further automation is to use

just the engine on a test bench, mathematical models and a controlled dynamicbrake to emulate the rest of the powertrain, the vehicle, the driver and the road.

This approach has been used in this work. In this way more reproducible mea¬

surements are obtained and comparisons become more reliable without the

necessity of time consuming hardware preparations.

If new concepts need to be analyzed a simpler tool than a complex emulation

can be deployed to deliver sufficiently accurate results. The quasi-static simu¬

lation of a vehicle during a driving cycle used to obtain the results in this chap¬ter is such a tool, that is able to estimate the fuel consumption of classical and

novel powertrain concepts with errors typically below 5%.

98 Chapter 8 A case study: MVEG-95 Fuel Consumption

8.1 Quasi-Static Simulation for Fuel ConsumptionEstimation

The quasi-static driving cycle simulation starts from the driving cycle veloc¬

ity and gear profile rather than from the engine torque. Thus, the torque and

speed required at the wheel to follow the velocity profile of the driving cycleis calculated depending on the friction and the inertia of the vehicle. With

models of each powertrain device the correct torque and speed demand to the

engine is derived allowing a table look-up approach to obtain the fuel con¬

sumption at every time step of the driving cycle. Effects such as for exampleshort variations of the fuel-air mixture due to flow dynamics or other similar

issues are not considered in this approach. The flowchart of the quasi-staticsimulation is given in Figure 8-1. One important restriction is that the driving

cycle simulation has no model implemented to consider cold start effects. As a

consequence all of the experiments were done with a warmed-up engine.

The New European Driving Cycle (MVEG-95) was used throughout all sim¬

ulations and experiments. It consists of 5 parts with the first one representing

driving in an urban area. This first part is repeated three times before the last

part, which emulates an extra-urban driving cycle with peak velocities of

120 km/h. For a vehicle equipped with a manual gear box, velocity and gear

are prescribed at every time interval, Figure 8-2. Further information on the

driving cycle can be found in [49].

Parameters used for the vehicle simulation are taken from the Ford Mondeo

model designation sheet and are listed in Table 8-1.

8.1 Quasi-Static Simulationfor Fuel Consumption Estimation 99

I Time J

Driving Cycle

distance

speed

m

Vehicle Model

vehicle torque and speed

T----gear ratio

-3.—

,

Transmission Model

2 c

CD t

> ©

o

Q_

clutch torque and speed

lPeripherals Model

engine torque and speed torque

Engine Model (Fuel Consump. Map)

Fuel Consumption

fuel flow

my= XmfAt

fuel massJ

Figure 8-1: Quasi-static simulation - flowchart.


7o40

-30CD

£20CD

Ö 10sz

CD

> 0A

o

Ä J\E K200 400 600 800 1000 1200

200 400 600 800

Time in Cycle [s]

1000 1200

Figure 8-2: New European Driving Cycle - MVEG-95.

Vehicle type: Sedan

Vehicle curb weight: 1325 kg

Vehicle length: 4481 mm

Vehicle width: 1749 mm

Tire: 195/60 R14

Air drag coefficient cd : 0.31

Frontal area A: 2.055 m2

Wheel radius rul :W

0.29 m

Rolling friction coefficient cr : 0.01 +1.6 10"7v3

Table 8-1: Main vehicle parameters.


8.1.1 Main Equations

Only the main equations of the simulation are given below. The sub-models

used are not explained here but can be found in [42], [44] and [47]. Without

considering wheel slip and assuming the vehicle on a flat surface the torqueneeded at the wheel Tw to follow the vehicle speed v as given by the driving

cycle is

r airÂ

2\_

' ~

w

Tw =

[m-8-cr+-^--cd-A-v -rw +

(A 0... ^+ m- rw

yW

V w jJt «»>

with m the total vehicle mass, cd the drag coefficient, A the frontal area of the

car, rw the wheel radius and 0w the wheel inertia.

The angular speed of the wheel depends only on the wheel radius and actual

vehicle speed

Given the above wheel torque and speed, the remaining torque and speed of

the other devices, that is the differential drive, final drive, gear drive, clutch

and engine, needs to be approximated (Figure 8-3). To simplify the simula¬

tion, the exact inertia of the wheels, brake disks, shafts, etc. are expressed as a

fraction of the total vehicle mass. For the simulation the vehicle mass is there¬

fore increased by 5%. The torque from the wheel will be increased bymechanical losses in each component, bearing in mind the backward calcula¬

tion approach of the simulation. Moreover this torque will be transformed bythe differential drive before being converted by the gearbox. The torque T

and speed coc after the clutch are therefore

T -i.

j, _

w t

C"

W' Tw> «J(8.3)

lt


y> Wheels

Required Engine Speedand Torque, thus Actual

Fuel Consumption.

Computation direction

Figure 8-3: Powertrain scheme.

Desired

Wheel Speedand Torque

with r\t the efficiency of the whole transmission depending on the gear ratio i,

the wheel torque 7^ and the speed co .The transmission gear ratio i. incor¬

porates the gear box and the differential drive in contrary to i, denoting the

gear ratio only.

The clutch model takes care that the engine speed (ùe ,when the engine

delivers power, does not drop below a certain speed threshold denoted as

co ..In that case, the clutch starts to open and the engine speed remains

t- fit ! ft

constant at co . without the torque being changed.c. fit 111


Wc for co > coc e, min

CO •(CO )

e, miny cJfor m < co .

c e, min

», = i. , ,

(8-4)

Finally the torque which the engine has to deliver is calculated as

Jco

Te = Tc + Taux + ®e-^f (8-5)

where 0g is the engine rotary inertia. Any auxiliary device is modelled by an

additional torque T which needs to be supplied by the engine. This torque

may depend on engine speed or triggering events at certain times during the

driving cycle.

With this set of equations the engine torque and speed can be calculated at

any moment during the driving cycle. The fuel consumption is obtained with a

two dimensional look-up table from the fuel map given in Chapter 5.2. Inter¬

nal engine friction needs no separate consideration since it is already included

in the fuel consumption map. Deceleration fuel cut-off is not implemented but

during braking the engine speed is reduced linearly instead of dropping imme¬

diately to idle. During gear changes, in contrast, the engine speed briefly goes

to idle level before increasing again after closing the clutch.

8.1.2 Validation

The regime in which the engine is operated during a driving cycle represents

approximately a quarter of the total engine operating range. Therefore onlythe fuel map for this part is needed for the simulation

.

The fuel consumption results of the driving cycle simulation can be com¬

pared with the experimental results. For better understanding the desired vehi¬

cle velocity is also shown in Figure 8-4 with a dashed line. The measured fuel

consumption is represented with a gray line whereas the black line is the fuel

consumption predicted by the simulation. As expected, the fuel flow during

gear change does not stop abruptly in a real engine but decreases rapidly

1. see Figure 5-1


Urban Cycle

Q.

E

COc

o

o

"CD

200

100 150

Extra-Urban Cycle

CDCDQ.

C/D

200 o

o

CD

>

0 100 200 300

Time in Cycle [s]

400

Figure 8-4: Instantaneous fuel consumption comparison.

toward the value at idle. As soon as the clutch is engaged the fuel consumptionincreases again. The simulation model has no dynamic effects implementedand thus cannot resolve this phenomena unless the engine speed is forced to

decrease not faster than a certain value. The fuel cut-off behavior duringdeceleration shows also some minor differences. Again a better model could

be used to improve the way the engine acts when the engine decelerate the

vehicle. As can be seen from the simulation, each time the vehicle starts from

standstill the fuel consumption increases and stays on a plateau for some

while. The reason for this plateau is the slipping clutch: during the first few

seconds, when the vehicle starts to move, the simulation asks for a lower

engine speed than the minimal allowed engine speed. The constant increase of

vehicle velocity results at a certain time in higher engine speeds than the min¬

imal allowed after which the clutch is closed completely, increasing fuel con¬

sumption.

8.1 Quasi-Static Simulation for Fuel Consumption Estimation 105

Another clearly distinguishable effect can be seen by analyzing the mea¬

sured fuel consumption just before the vehicle starts to accelerate after stand¬

still. Since a predictive driver was implemented in the test bench emulation,

the engine speed is raised before the vehicle starts to move to assure a good

matching of the desired velocity trace. Therefore the fuel consumption rises

earlier than in the quasi-static simulation.

The extra-urban driving cycle (EUDC) shows an almost complete agreement

between the simulations and the measurements since gears are not shifted as

often during the EUDC as during the urban cycle.

The cumulated fuel consumption is given in Figure 8-5. The two curves can

hardly be distinguished from each other. At the end the overall fuel consump¬

tion differs by just 0.33%. The absolute values are 0.89476 kg for the simula¬

tion and 0.89733 kg for the measurement. To better interpret those values, the

fuel consumption error during the cycle is shown in Figure 8-5. It was calcu¬

lated with the following formula.

Fuel consumption error -

> mf At—ymr At/ j j, sim Z^ j, meas

1m_e Atf meas

(8.6)

As can be seen, the error increases when gear changes are performed and

decreases during decelerations. These two conflicting errors tend to average

out and are always well below 1%.

Another important aspect is the percentile fuel relevance R r of an operating

point. The real engine and the simulated engine should follow ideally the same

time depending trace in the fuel map. In order to quantify the relevance of an

operating point, the fuel map was divided into individual regions of 5 Nm

times 200 rpm each. The fuel consumption of this region was set to the mean

value of the fuel consumption of this area.


8-o50.

co

|o.</3

C

o

O

©0,

LL

^

o

I-0.2 h

0 1

QSS««««"» Measured

0 200 400 600 800

Time in Cycle [s]

1000

Figure 8-5: Comparison of cumulative fuel consumption.

200

0.8

^ 0.6

0.4

LU

0?(_

o-t—«

Q. 0e

-0.2

o

O-0.4

CD13

u_ -0.6

-0.8

200 400 600 800

Time in Cycle [s]

1000 1200

Figure 8-6: Continuous fuel consumption error.


The relevance of certain areas on the total fuel consumption are calculated

using the following relation.

£*mpAt

RfTe r.T—,coe «• +

—

100 (8.7)

T* ffli XmtAt

The result can be seen in Figure 8-7 and Figure 8-8. Only the values of at

least one percent of the total fuel consumption were drawn and therefore the

the fuel relevances of all visible fields do not sum up to 100%. It is significantthat idle is relevant with as much as 9% in the simulation and 6% in the mea¬

surements. Thus cutting off fuel instead of idling and restarting the engine as

soon as power is needed again, results in significant fuel savings as will be

seen in Chapter 8.2.1.

Finally it should be noted that a major benefit of using the vehicle emulation

on a test bench and the quasi-static simulation is that the parameter file con¬

tains equal system values, as for example tire friction or vehicle drag. The

results are therefore comparable and show that a quasi-static driving cyclesimulation delivers accurate results. For this reason it is assumed to be a suit¬

able tool to estimate fuel improvements whenever a new powertrain is pro¬

posed.

Consequently it is acceptable to use the quasi-static simulation for the evalu¬

ation of the fuel consumption reduction expected during a driving cycle if sec¬

ondary valves are used to control the engine torque.


80

70

60

— 50CD

CT

o40\-

CD

5,30c

LU

20

10

0

i

T 1 !Tl

1

1 2

3 2

2 2 !3 2

2

41 1^

1

9

j 3

3 6

23 ;3 2L_ i

0 1000 2000 3000

Engine Speed [rpm]4000

Figure 8-7: Fuel relevance during MVEG-95 - measurement.

80

70

_60E

^ 50CD

CT

o40I-

CD

1,30c

LU

20

10

0

: 2 !

: 2

2 :

3

3 3

1

1 3 ;2 1

1 : 6

W"~"11

e 5

i\ 8

9 2i i

1000 2000 3000

Engine Speed [rpm]

4000

Figure 8-8: Fuel relevance during MVEG-95 - simulation.

8.2 Comparison ofExperiments and Simulations 109

8.2 Comparison of Experiments and Simulations

By using the fuel maps of the conventional and secondary valve engine of

the previous chapters four different engine configurations were tested usingthe driving cycle simulation.

1. conventional engine2. secondary valve engine3. secondary valve engine with reduced idle fuel consumption4. conventional engine with automatic stop and cranking device

The conventional throttled engine served as base for all comparisons and

with the evaluation of the secondary valve engine the final conclusion of this

work is drawn. Since the idle fuel consumption has a major impact on the

overall fuel consumption two other simulations were performed. This allows

to compare secondary valves with other ways of reducing the overall fuel con¬

sumption. Therefore, in the secondary valve case, the idle fuel consumptionwas reduced linearly in function of speed as discussed in Chapter 3.1.3. In the

conventional throttled case an automatic stop and cranking device was imple¬mented. Usually, at engine start some extra fuel is injected to ensure a smooth

and immediate running but this effect was not considered in the driving cyclesimulation nor that operating the cranking device needs some energy as well.

8.2.1 Driving Cycle Simulation - Results

The results of all four simulations are shown in Figure 8-9. The total fuel

consumption of the conventional engine is 0.895 kg compared to 0.839 kg in

the secondary valve case, a fuel consumption improvement of 6.2%. Up to

7.8% of the fuel can be saved, if the idle speed fuel consumption is reduced by

using the secondary valves by means of lowering idle speed, equaling0.825 kg of fuel. A stop and go device was simulated as well to allow an

assessment of an engine equipped with secondary valves. Such a device per¬

mits highest improvements reducing the fuel consumption to as low as

0.764 kg or an improvement of 14.6% compared to the base measurement.

Automatic cranking can also be deployed on a secondary valve equipped

engine. Since in the secondary valve case the idle fuel consumption remains


the same as in the conventional throttled engine, the fuel improvements of the

automatic cranking device can directly be counted towards the fuel improve¬ments of the secondary valve engine. This would lead to a fuel consumption of

0.708 kg fuel, or to 20.9% fuel saving respectively.

Figure 8-9: Final comparison of cumulative fuel consumption.

Ill

Chapter 9

Conclusions and Outlook

By experiments and simulations, it was proven that secondary valves are

capable of operating in both the anticipated regimes - load control by using

early effective intake valve closure and roughness control by reducing the

effective valve overlap of intake and exhaust valves.

Throttleless engine charge control with an early valve closing scheme signif¬

icantly reduces pumping losses but deteriorates the burn rate because of a

sharp temperature and pressure drop after the secondary valve closed. Espe¬cially at low load, when pumping losses are considerably high and therefore

their reduction desirable, a secondary valve load control scheme shows its

weakness. This is not due to the secondary valve system itself but is an

unavoidable consequence of the early valve closing scheme - any similar load

control approach has to overcome the drawback of burn rate deterioration. For

example at an engine speed of 1000 rpm the fuel benefits decrease from 9.1%

at 18 Nm load to only 6.1% at 8 Nm load, which is a reduction of 33%. Due to

the burn rate deterioration the secondary valve closing scheme shows best effi¬

ciency at approximately 32 Nm decreasing slightly to lower loads at higher

engine speed.

Engine roughness control with secondary valves is particularly promising.Camshafts with large valve overlap can be used at lowest load without increas¬

ing the roughness much higher than the level of an engine with conventional

camshafts. This permits a significant rise of peak power of at least 10% due to

improved of the engine at higher speed. At 980 rpm and 20 Nm load a rough-

112 Chapter 9 Conclusions and Outlook

ness reduction of 80% was achieved with secondary valves. To reach the

roughness of a conventional camshaft equipped engine a reduction of 100% is

required. At 1600 rpm, it was possible to reduce engine roughness to the level

of the conventional engine.

A very fast secondary valve actuator is able to control load and roughness at

the same time, because load control uses the closing phase of the secondaryvalve and roughness control the opening phase. In the case the secondaryvalve actuator does not permit such fast operation (< 1 ms) it should be noted

that using secondary valves as load control device already reduces engine

roughness partially since manifold pressure remains almost atmospheric. This

leads to reduced residual gas backflow, improving considerably burn rates.

The conclusions drawn from the engine simulations are that load control

with secondary valves, despite the drawback mentioned above, reduces the

fuel consumption of an engine and that the parasitic volume does not alter the

benefits significantly. Taking into account the energy consumption of the sec¬

ondary valve actuators the fuel benefits are significantly reduced or may even

vanish. On this basis, it can be deduced that fuel consumption benefits

obtained when controlling engine load with secondary valves does not justifythe equipment necessary for its successful deployment. Nevertheless takingfull advantages of secondary valves by implementing different operationmodes simultaneously such as load control, roughness control, cylinder indi¬

vidual control and cylinder cut-off, could make further investigations interest¬

ing.

The second aim of this thesis, estimating the fuel consumption of a vehicle

during a driving cycle by using two simulations, one for the engine cycle and

one for the vehicle, was successfully demonstrated. It is now possible to eval¬

uate a new engine component regarding fuel consumption acting on the ther¬

modynamic engine process. Optimizing engine and vehicle as a whole should

therefore be facilitated, allowing improvements of the entire system instead of

improving each part separately.

113

With these tools it was estimated that a mid-class vehicle equipped with a

2 liter engine uses during a MVEG-95 driving cycle 0.895 kg of fuel. The use

of secondary valves reduces this fuel consumption by 0.056 kg to 0.839 kg, a

6.2% improvement.

A broad range of further investigations are conceivable. Up to now, the sec¬

ondary valve scheme was only used at stationary engine operating points.

Consequently the next step should be the design and implementation of

improved secondary valves permitting transients to be carried out and thus

permitting driving cycle experiments on a test bench. For this, a more reliable

and higher performance actuator with a sophisticated secondary valve control¬

ler is needed. This controller could use the secondary valve timing map

derived in this thesis as a first starting point.

Another possible area of investigation is cylinder cut-off, which could be

easily implemented if driving cycles are already realized using an engine

equipped with secondary valves. Unfortunately load control and cylinder cut¬

off fuel benefits cannot be added, nevertheless the remaining cylinders will

still be controlled by secondary valves. The increased load in these cylinderreduces the burn rate deterioration and allows increased secondary valve per¬

formance.

Finally, the optimal combined use of secondary valve and throttle could also

be analyzed. In fact, a secondary valve most probably will not be able to oper¬

ate at all engine speeds, the right point for secondary valve load control to take

over from the conventional throttle has therefore to be determined.

115

Appendix A

Parameters

A.l Main values - engine

Engine:

Emission norm:

Fuel:

Engine code:

Firing:

Bore:

Stroke:

Clearance volume

Displacement:

Compression ratio:

Power (DIN):

Torque (DIN):

Idle speed:

Maximum speed:

Ford Zetec 2.0 1 DOHC 16 V

83 US/93/59 EEC

Super 95 ROZ, unleaded

NGA

1-3-4-2

84.8 mm

88.0 mm

55.2 cm3

1988 cm3

10.0:1

100 kW at 6000 rpm

180 Nm at 4000 rpm

800 rpm

7100 rpm

2 Mam values - vehicle model

Vehicle Type

Vehicle length:

Vehicle width:

Air density:

Air drag coefficient:

Vehicle curb weight:

Tire:

Wheel radius:

Rolling friction coef]

Frontal vehicle area:

1. gear

2. gear

3. gear

4. gear

5. gear

Final drive

Ford Mondeo Sedan

4481 mm

1749 mm

1.15 kg/m3

0.31

1325 kg

195/60 R14

0.29 m

cr- 0.01 + 1.6 10"7v3

2.055 m2

3.231

2.136

1.438

1.114

0.854

4.06

117

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"I Lw L681 /

J!ank lea

Curriculum Vitae

Name:

First names:

Date of birth:

Place of birth:

Vogel

Olivier Denis

3. January 1969

Thalwil, Switzerland

Education

1976-1982

1982-1985

1985-1989

1989-1995

1995

1995-2000

Primary school, Adliswil

Secondary school, Adliswil

High school, Zurich

Studies in mechanical engineering,Swiss Federal Institute of Technology (ETH), Zurich

Diploma in mechanical engineeringSwiss Federal Institute of Technology (ETH), Zurich

Doctoral student and research assistant at the

Institute of Energy Technology,Swiss Federal Institute of Technology (ETH), Zurich

in copyright - non-commercial use permitted rights ... · ix abstract this thesis analyzes a novel...

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