VARIABLE VALVE TIMING (H ÖGRE · PDF fileoptimized flex-fuel engine with fully variable...

OPTIMIZED FLEX-FUEL ENGINE WITH FULLY

VARIABLE VALVE TIMING (HÖGRE

OTTOMOTORVERKNINGSGRAD MED NY

VENTILAKTUATOR VILKET GER REDUCERING

AV CO2 OCH BRÄNSLEFÖRBRUKNING)

FEASIBILITY AND INVESTIGATORY STUDY

FINAL REPORT

AVL SPEAB

JOHANNES ANDERSEN

PER-ERIK NILSSON

JOAKIM KARLSSON

ÖYVIND GUNDERSEN

SÖDERTÄLJE POWERTRAIN ENGINEERING AB

2011/11

A REPORT FOR

THE SWEDISH ENERGY AGENCY

SAAB POWERTRAIN AB

CARGINE ENGINEERING AB

Document - Type Final report

Prepared by Johannes Andersen

Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4

Document - Ref Optimized Flexfuel Engine with FVVT_20111130

Page 2 (70)



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AVL SPEAB Transmissionsvägen 2 SE-151 48 Södertälje Sweden Phone: +46 8 500 656 00 Fax: +46 8 500 283 28 e-mail: [email protected] Web: http://www.avl.com/



Per-Erik Nilsson

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ABSTRACT A pneumatic fully variable valve timing system designed by Cargine has been adapted to a spark ignited single cylinder research engine at AVL and tested to evaluate the potential fuel consumption reduction achievable with this specific technology. Valve strategies, fuels, fuel injection method and compression ratio as well as traditional operating parameters such as spark advance, start of injection etc. have been varied during the tests. The results cannot be judged solely on the basis of fuel consumption reduction nor efficiency; it is imperative to distinguish and separate what can be achieved with the fully flexible valve train in terms of reduction of pumping losses and how this particular valve system interacts with the combustion system that has been used for this investigation.

Some of the conclusions drawn from the project:

• Fuel consumption reductions achieved in comparison with engine using continuously variable cam phasing are:

o 0,2 – 10,9% depending on operating point, with an average of approximately 6% for same compression ratio

o 6,8 – 17,5% depending on operating point, with an average of approximately 12% when comparing a 13:1 compression ratio fully variable valve timing engine with a 9,2:1 compression ratio camshaft reference engine

• It has been demonstrated that the fully variable valve timing can facilitate an increase in compression ratio by approximately 4 units while still achieving comparable full load engine torque of a camshaft engine

• Some of the part load operating points do not show the expected benefits when using the fully variable valve timing even though the pumping loss reduction is clearly evident in the cylinder pressure trace

• For supercharged operating points the fully variable valve timing enable higher performance and efficiency for the same boost level

• It is possible to idle the engine without throttling, however at severe pumping losses

• The parasitic losses of the fully variable valve system when using 4 valves is comparable to a standard camshaft setup at lower speeds (up to 2000 rpm), while at 3000 rpm the increase in parasitic losses is approximately 10%

• It has been demonstrated that the fully variable valve system has acceptable stability within the tested speed and load range



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SAMMANFATTNING Ett pneumatiskt system för fullt variabla ventiltider konstruerat av Cargine har anpassats till en encylindrig forskningsmotor av otto-typ vid AVL och utprovats för utvärdering av bränsleförbrukningsvinster. Olika ventilstrategier, bränslen, insprutningsmetoder och kompressionsförhållanden samt även konventionella motorparametrar såsom tändvinkel, insprutningsvinkel etc. har varierats under proverna. Resultaten kan inte enbart bedömas utifrån erhållen bränsleförbrukning och verkningsgrad; det är viktigt att separera vad som kan åstadkommas med hjälp av de fria ventiltiderna i form av reduktion av pumpförluster med hur detta ventilsystem interagerar med det aktuella förbränningssystemet som använts under undersökningarna.

Ett urval av resultaten sammanfattas nedan:

• Uppmätt bränslebesparing i jämförelse med en motor med vridbara kamaxlar är:

o 0,2 – 10,9% med ett medelvärde på 6% för samma kompressionsförhållade

o 6,8 – 17,5% med ett medelvärde på 12% i en jämförelse mellan en motor med fria ventiler samt förhöjt kompressionsförhållande (13,1:1) och en motor med vridbara kamaxlar samt standard kompressionsförhållande (9,2:1)

• Prover har demonstrerat att för en motor med fria ventiler kan kompressionsförhållandet kan höjas med 4 enheter och ändå uppnå likvärdig prestanda som en motor med kamaxlar

• Vid ett antal dellastpunkter uppmättes inte förväntad bränsle-förbrukningsreducering trots att en reduktion i pumpförluster tydligt framgår i det indikerade cylindertrycket

• Fria ventiler möjliggör vid överladdning en högre effekt och verkningsgrad vid samma laddtryck jämfört med ett kamaxelsystem

• Otrottlad tomgång är möjlig med det fria ventilsystemet, dock endast med påtaglig pumpförlust

• Parasitförlusterna för det fria ventilsystemet är motsvarande de hos ett konventionellt kamaxelsystem vid låga varvtal(upp till 2000 rpm). Dock ökar skillnaden till ca 10% vid 3000 rpm

• Prover har demonstrerat att de fria ventilerna har acceptabel stabilitet vid de laster och varvtal som utprovats inom projektet



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TABLE OF CONTENT

1 INTRODUCTION ............................................................................................................................ 8

1.1 BACKGROUND .......................................................................................................................... 8 1.2 PROJECT DEFINITION ................................................................................................................ 9

1.2.1 Targets ............................................................................................................................... 9 1.2.2 Planning ............................................................................................................................. 9 1.2.3 Stakeholders .................................................................................................................... 10 1.2.4 Execution ......................................................................................................................... 10

1.3 BASIC PRINCIPLE OF ACTUATORS TESTED ................................................................................ 14 1.3.1 FVVT generation 1 ........................................................................................................... 14

2 INFORMATION COLLECTED ..................................................................................................... 17

2.1 VARIABLE VALVETRAINS IN USE TODAY ..................................................................................... 17 2.1.1 Honda V-Tec .................................................................................................................... 17 2.1.2 BMW Valvetronic ............................................................................................................. 18 2.1.3 Fiat MultiAir ...................................................................................................................... 19 2.1.4 Lotus AVT ........................................................................................................................ 20 2.1.5 Cargine Free Valve Technology ...................................................................................... 21 2.1.6 System comparison ......................................................................................................... 21

2.2 LITERATURE SURVEY .............................................................................................................. 22

3 DESIGN AND IMPLEMENTATION .............................................................................................. 28

3.1 ENGINE SPECIFICATION ........................................................................................................... 28 3.2 FVVT INSTALLATION ............................................................................................................... 28

4 EXPERIMENTAL SETUP ............................................................................................................. 29

4.1 INSTRUMENTATION ................................................................................................................. 29 4.2 CALCULATIONS AND ASSUMPTIONS .......................................................................................... 31

4.2.1 Compressor power .......................................................................................................... 31 4.3 FUELS .................................................................................................................................... 33

4.3.1 E10 ................................................................................................................................... 33 4.3.2 E85 ................................................................................................................................... 34

4.4 TEST PROCEDURE .................................................................................................................. 35 4.4.1 General ............................................................................................................................ 35 4.4.2 Part load ........................................................................................................................... 37 4.4.3 Full load ........................................................................................................................... 37 4.4.4 Idle ................................................................................................................................... 38 4.4.5 Parasitic loss investigation ............................................................................................... 38 4.4.6 Remarks to test procedure .............................................................................................. 39

5 EXPERIMENTS ............................................................................................................................ 40

5.1 PART LOAD OPERATION ........................................................................................................... 40 5.1.1 Summary .......................................................................................................................... 40 5.1.2 Detailed data .................................................................................................................... 44 5.1.3 Tests performed with port fuel injection, PFI ................................................................... 48

5.2 FULL LOAD OPERATION ........................................................................................................... 49 5.2.1 Summary .......................................................................................................................... 49 5.2.2 Naturally aspirated ........................................................................................................... 49 5.2.3 Supercharged .................................................................................................................. 53



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5.3 IDLING .................................................................................................................................... 58 5.3.1 General ............................................................................................................................ 58

5.4 OVERALL ANALYSIS................................................................................................................. 61 5.4.1 Power consumption ......................................................................................................... 61 5.4.2 Reference engine versus high compression FVVT engine ............................................. 63 5.4.3 Valve stability effects ....................................................................................................... 64 5.4.4 Direct injection versus port fuel injection and other combustion system related discussions ................................................................................................................................... 64 5.4.5 Higher engine speed and FVVT ...................................................................................... 64

6 CONCLUSIONS ........................................................................................................................... 65

7 RECOMMENDATIONS AND OUTLOOK .................................................................................... 67

8 REFERENCES ............................................................................................................................. 68

8.1 WEB ...................................................................................................................................... 68 8.2 LITERATURE ........................................................................................................................... 68

9 NOMENCLATURE ....................................................................................................................... 69



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Date – Rev 20111130-4


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1 INTRODUCTION

1.1 Background

Combustion engines of today have evolved on the basis of the customer requirements on drivability, price, fuel economy and of course the legislative emissions. In future engines and in particular the spark ignited, SI, engine, the demands on fuel economy will increase substantially. The European Automobile Manufacturers Association, ACEA, agreement of 140 g/km of CO2 by 2008 (not met) and 130 g/km of CO2 by 2015 [1] as well as the European Union binding target of 20% renewable energy and 20% energy efficiency of 2020 [2] clearly sets these demands. In the US similar requirements are applied, such as the Corporate Average Fuel Economy, CAFE, standard. Diesel engines are not only subject to these demands but also currently undergoing a tough adaptation to new emission levels which in combination with the fuel economy targets clearly sets new challenges in the design of modern combustion engines.

In order to meet the new targets and still produce a financially viable product it is necessary to develop new technology that will increase efficiency and also lower the emission levels. The difficulties have typically been related to the varying working conditions that an automotive engine will experience, i.e. the engine is lacking the flexibility to maintain a high efficiency and low emission level throughout the whole working range given the fixed geometries that has been the results of various compromises during the development.

Putting the combustion aside it is clear that a combustion engine to a large extent is pumping the working media in and out of the cylinder. For different speeds and different requirements of mass flow running through the engine, the engine should ideally act as a variable pump that does not rely on throttling inlet or outlet to reach the correct mass flow. This applies to the SI-engine in particular since the engine is typically running stoichiometric air/fuel ratio, AFR, and therefore performance is controlled by the amount of air going into the engine using a throttle. A higher degree of freedom for the valve timings of the engine is necessary in order to minimize the throttling for the SI-engine. In the case of a diesel engine the load is basically controlled by the amount of fuel injected, but even in the diesel engine the gas exchange process, when dependent of camshafts with fixed durations, is always a compromise between different operating points.

Ideally, the valve timings should be adjusted so that the engine can achieve the desired gas exchange process without unnecessary compromises even though the speed and load changes. For both the diesel engine and the SI-engine the use of a fully variable valve train, FVVT, would increase the efficiency of the gas exchange dramatically. Another potential benefit could be more exotic



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combustion concepts that can be realized with FVVT such as skip cycle/8 stroke as well as two-stroke modes.

The combination of alternative fuels and the flexibility that FVVT offers is particularly interesting, as previous optimizations and compromises have to a rather large extent been determined by fuel properties. For example, the possibility of manipulating the geometrically based compression ratio with valve timings to adapt the engine to different fuel qualities is typically one area of interest.

1.2 Project definition

1.2.1 Targets

The project is intended to investigate and quantify the benefits of the pneumatically driven FVVT designed by Cargine Engineering on a single cylinder SI-engine running steady state. In short the project will:

• Investigate the behavior and feasibility of the pneumatic FVVT design

• Quantify the increase of engine efficiency and fuel economy using the FVVT

• Quantify the parasitic losses connected to running the FVVT

• Quantify the reduction in emissions when using the FVVT

• Investigate new combustion concepts that are possible using FVVT

• Investigate and quantify any synergies when using alternative, renewable, fuels together with the FVVT

1.2.2 Planning

The project is divided into three parts:

• Design, i.e. adapting the actuators to the engine

• Development of an in-house control system to facilitate control of the actuator solenoids and other engine related signals such as spark timing, injection parameters etc.

• Testing and evaluation

The project was approved for governmental funding by the Swedish Energy Agency 2010-08-18, with retroactive funding back to 2009-12-19. Substantial preparation work had however been in progress well before the time of approval. The project end date is 2011-09-15.

In this case the project has also involved the work of two master theses, the conclusions of which are also summarized in this report.



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1.2.3 Stakeholders

The involved stakeholders and their associated responsibility are presented in Table 1.

Table 1, Stakeholders and responsibilities

Stakeholder Responsibility Roles/resources

Cargine AB Design of actuators, problem

solving related to actuator

function

Designer, test engineer

AVL SPEAB Design and manufacturing of

cylinder head with the necessary

adaptation for the FVVT

Designer, mechanic

SAAB Powertrain Steering committee input Senior Engineer

AVL SPEAB Testing in single cylinder dyno Test engineers, EMS

engineers, dyno

technicians, mechanics

and test cell equipment

AVL SPEAB EMS development EMS engineers, HW

modules, licenses etc.

Swedish Energy

Agency

Funding, approval of final report -

Finnveden powertrain AB was initially a stakeholder in the project, however at the start of the project Finnveden withdrew their involvement due to lack of resources.

1.2.4 Execution

The project is based on tests performed in a single cylinder test bed. The single cylinder test bed is ideal for investigations relating to combustion properties and is easily adaptable for different engine concepts. The test bed incorporates features such as the possibility to run the engine supercharged as well as multi-fuel capability. Four actuators have been installed in a SAAB cylinder head meaning that both inlet valves and exhaust valves are controlled by the actuators. The engine used has a wall guided direct injection system for introducing the fuel directly into the combustion chamber. The engine is normally equipped with continuously variable cam phasers, CVCP, that can advance or delay the valve lift curve defined by the camshafts. During the tests the engine displacement has been 0,55 liter instead of the 0,50 liter of the standard engine.



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The pneumatic actuator is developed by Cargine Engineering AB and the adaption to the cylinder head has been made jointly by AVL and Cargine. The first design of the actuator had a box-like design and used two solenoids for control, see Figure 1.

Figure 1, Valve actuator with position sensor

The tests include:

• Part load operation at points that are derived from the New European Drive Cycle, NEDC

• Idling

• Performance when running naturally aspirated

• Performance for supercharged engine

The parameters that were altered between and during the tests were typically valve timings, valve lift, different fuels, geometric compression ratio, start of injection, injection duration, spark advance, etc.

In particular, the manipulating of compression/expansion ratio using the Atkinson cycle is of special interest, as is the possibility of very short inlet valve durations often referred to as the Miller cycle. Both of these modes can be used at low engine loads, and the Atkinson cycle can also be used at high loads given sufficient amount of boost pressure being available. See Figure 2 for schematic pV diagrams of the cycles.



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Figure 2, The Atkinson and Miller cycle

Fuel consumption reduction up to 11% has been shown using the Atkinson cycle [4] and 3% increased brake efficiency has been proven on turbocharged diesel engines [5].

Skip cycle tests in which the engine, for each cylinder, does not fire once every second revolution, but instead once per every four revolutions is also possible with FVVT. The inlet valve event is performed as usual but the engine then turns two revolutions with the valves closed after which the ignition occurs, followed by an opening of the exhaust valves and the cycle is then completed. See Figure 3 for a motored pressure and valve lift trace without combustion.



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-360 0 360 720 1080 1440 1800 2160 2520 2880 3240 3600Crank Angle [deg]

Val

ve li

ft [m

m]

-3

0

3

6

9

12

15

18

PC

YL

1 [b

ar]

-2

0

2

4

6

8

10

12

14

16

18

Figure 3, 8-stroke pressure and valve trace, motored with no combustion

The idea is to have lower losses due to the engine is running at a higher load for the combustions that are taking place for the same overall power output. An alternative to the sequence above would be to place the combustion at first compression, empty the cylinder of exhausts and recompress 2 times before opening the inlet valve.

In many ways similar to skipping cycles for each cylinder, the strategy of cylinder deactivation may also be realized with the FVVT. With the valves closed during the deactivation, flow losses are minimized even though the heat losses would still be present. As this project is based on tests with a single cylinder test bed the limiting impact of engine harshness could not be determined, i.e. this mode needs to be evaluated on a full sized engine.

FVVT can also be used to run the engine at 2-stroke mode, this could potentially be realized at high load and lower engine speeds at which the positive pressure differential between inlet and exhaust ports enables a simultaneous opening of inlet and exhaust valve and hence promote scavenging. The internal EGR can nevertheless be expected to be high, and while the direct injection might result in minimal amounts of fuel being directly lost through the exhaust valve the time for the fuel preparation versus spray collision with the piston is heavily reduced. The positive result of 2-stroke mode

Recompression/re-expansion

Combustion would occur here

Valve events

Normal 4-stroke cycle

8-stroke cycle



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is naturally the possibility of twice the amount of power strokes compared to the 4-stroke engine and potentially twice the power from the same displacement.

The parasitic losses related to the air consumption of the pneumatic FVVT are calculated according to concept that is based on an installation in which no cooling of the heated compressed air takes place and in which the pressurized air is reused without fully expanding to ambient pressure. Schematic and details of the calculations are presented under section 4.2.

The engine management system, EMS, has been developed in several stages throughout the project. Several closed loop modes are available and there has also been a build-up of model based control modes that in a later phase could be used for continuous fuel consumption optimization. Further developments would also include algorithms handling switching between different modes, which has not been introduced for this project as the tests are solely based on steady state operating points.

1.3 Basic principle of actuators tested

1.3.1 FVVT generation 1

Each actuator, see Figure 4– left picture, controls one engine valve each and consists of an actuator piston, AP, seen in Figure 4 –right picture, cylinder, two solenoids, two spool valves, two port valves and a hydraulic latch. Solenoid 1, called the timing solenoid, TS, controls a spool valve and the hydraulic latch. Solenoid 2, called the lift solenoid, LS, controls another spool valve. In turn, the spool valves controls the air entering the actuator cylinder. For more details, see [6] and [7].

Figure 4, Free valve actuator with piston

The compressed air powers the piston which presses on the engine poppet valve, causing it to open, see Figure 5. To better describe the dynamics of the



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system, the valve opening and closing process can be divided into three parts; air charging, expansion and dwell and air discharging.

Figure 5, Actuator detail

1.3.1.1 Air charging state When the TS is energized, it opens spool valve 1 which then sends compressed air into the actuator cylinder. As the pressure rises, it pushes the AP outwards and it in turn opens the engine poppet valve, see Figure 5. An oil check valve is activated simultaneously with the TS, becoming a one way valve, stopping the oil flowing back to the reservoir.

All electrical solenoid signals used to control the oil and air flow in and out of the actuators are approximately constant in time, which means that there is a strong dependence on engine speed. The engine speed dependence imposes certain limitations on the valve variability:

• As engine speed increases, low valve lifts are limited by a minimum lift control time and if lower lifts are to be obtained the pressure feed must be lowered

• For higher engine speeds, the entire valve lift profile will change appearance drastically, Figure 6, which of course also affects the gas exchange behavior and efficiency



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Figure 6, Intake and exhaust valve lift at 800 rpm (left) and 3000 rpm (right)

Requested timings and lift heights are identical, but the curves look very different.

1.3.1.2 Expansion and dwell stage When the LS is energized, it opens spool valve 2 which stops the inlet air. However, since the TS is not yet de-energized, the hydraulic latch still allows flow into the actuator and the air can expand further. It completely expands until it equals the spring force and other resistance factors. The hydraulic latch secures the valve in the maximum lift height, preventing oil flowing back to the reservoir and holding the piston in place. A small oscillating movement can be seen as the valve compresses the oil. Using this strategy, to allow the compressed air to expand fully inside the actuator cylinder, it extracts full expansion work from the air and hence lowers the energy consumption. The timing between energizing the two solenoids, called the lift control time, is what decides the lift height and it is crucial to be able to control this carefully to be able to choose height with a reasonable degree of precision.

1.3.1.3 Air discharging stage When TS becomes de-energized, spool valve 1 returns to zero and the hydraulic latch is opened again. The air trapped in the actuator cylinder begins to discharge and the oil can flow freely back into the reservoir. Once the AP moves close to the valve seat the oil is blocked from the return line to the reservoir and is trapped, dampening the valve movement efficiently. Changing the distance, i.e. the dead volume distance, in the actuator, the point where the piston enters the braking zone can be determined and subsequently lowering the seating velocity.

1.3.1.4 Updated valve actuator At the end of the project an updated valve actuator was tested. The intention behind the update was primarily to improve the stability of the valve lift. The



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position sensor for valve lift has been kept for the evaluation of the valve behavior, however there is an additional switching sensor that has been added into the design of the actuator that is indicating 1 mm and 9 mm lift. The intention is that the switching sensor will be used for control purposes in the production actuator. Data from tests with this design will be clearly marked in the report.

2 INFORMATION COLLECTED

2.1 Variable valvetrains in use today

In this section summaries based on previous work [8] and [9] are presented.

A number of various valvetrain systems exist today. The most common arrangement in today’s automotive industry is to use one camshaft for the intake valves and one camshaft for the exhaust valves. Car makers that want to offer higher levels of performance implement some type of cam phasing, which means that the entire valve lift profiles can be advanced or retarded relative to each other and the crankshaft. The mechanical constraints typically limit the variation range to around 50 crank angle degrees, CAD. While the number of valvetrain systems capable of phasing the lift curves steadily increases, very few systems currently in production are able to handle variable lift heights. There are however some exceptions and a couple of them are listed below.

2.1.1 Honda V-Tec

Hondas V-Tec system (Variable valve Timing and lift Electronic Control) was introduced in the late 1980’s and essentially consists of a camshaft with two different cam lobes, see Figure 7. During low and moderate engine speeds and loads, the valves are operated with relatively short opening durations and low lifts. As speed and load increases, the entire camshaft is moved, thereby causing the cam lobe with longer duration and higher lift to actuate the valve. It is advantageous to link duration and valve lift to each other in order to reduce the mechanical load on the valves, i.e. it could be difficult to combine short durations and high lifts with this mechanical installation. The switching between the two modes is done via a hydraulic unit which is electronically controlled through a solenoid.



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Figure 7, Sketch of a V-Tec valvetrain system, the solenoid control strategy is based on engine speed and load

Various development modifications of this configuration have been introduced over the years, including cam shafts with three different cam lobe profiles.

2.1.2 BMW Valvetronic

The BMW valvetrain system was introduced in 2001 for a handful of different engine configurations. The system is able to continuously vary both intake valve lift height and phasing and is especially useful since it enables early inlet valve closing, IVC, and thus reduced pumping losses. Furthermore, the BMW engines can operate almost completely unthrottled but the throttle plate is actually kept since it is used during cold starts and in limp-home safety mode. To make the Valvetronic system work, the cylinder head is equipped with an extra set of rocker arms between the camshaft and the valve stem, see Figure 8. These rocker arms can, thanks to an electronically controlled additional camshaft, pivot freely. The mechanism makes sure that valve duration is a function of lift height, thereby avoiding excessive and possibly harmful valve accelerations and seating velocities. This lift mechanism is combined with a phaser, which makes the choice of inlet valve opening, IVO, and IVC almost independent of each



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other. The individual intake valve lifts can also be offset relative to each other in order to increase air charge motion even more around idling speed.

Figure 8, The Valvetronic system, the movements of the original camshaft and the eccentric camshaft are combined to generate continuously variable valve lifts

Since the Valvetronic system requires stronger valve springs than the same engine configuration with fixed cam shafts, a small efficiency drop at higher engine speeds will occur. Because of this, the high performance (high speed) BMW models currently run without the Valvetronic system.

2.1.3 Fiat MultiAir

Fiat introduced the MultiAir system in 2009 and it has currently been implemented on a couple of vehicle models. A MultiAir engine is based on a traditional mechanical exhaust camshaft. This is connected to the intake valves via a hydraulic unit, see Figure 9. Since the oil is practically incompressible it works as a rigid element, thereby causing the intake valves to be synchronized with the exhaust valves. Through the use of an electronically controlled solenoid valve, however, the oil volume can be regulated and by timing this solenoid



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valve opening the intake valve timings become decoupled from the exhaust camshaft. The solenoid control is also sufficiently fast to enable multiple valve lifts, which could be used to induce turbulence at lower engine loads.

Figure 9, MultiAir valvetrain installation; the single camshaft is mechanically connected to the exhaust valves and the hydraulic unit incorporates the intake valves into the assembly

2.1.4 Lotus AVT

The Lotus AVT (Active Valve Train research system) is a fully variable valvetrain system that is primarily used for research and development purposes rather than mass production. Instead of camshafts, hydraulically operated actuators and electro-hydraulic servo valves are installed above the valves. It is also equipped with position transducers, which makes the entire valvetrain



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assembly very similar to the one developed by Cargine. The major difference is the use of hydraulics instead of pneumatics to control the actuator and valve motion.

2.1.5 Cargine Free Valve Technology

The valvetrain system used in this project comes from Cargine Engineering and is more thoroughly described in section 1.3. The pneumatics used to actuate the valves imposes certain limitations on the variability range, which cannot be seen in e.g. the Lotus system. Based on feedback from position sensors fitted to the valves, independent variation of lift height, opening and closing of each valve is achieved. The system behavior is engine speed dependent since all control signals are approximately constant in time. The control signal implementation also limits the lift height variation range to a certain extent; operation with valve lifts below 3 mm is difficult to maintain due to valve timing stability concerns.

2.1.6 System comparison

These systems all have their own specific advantages and disadvantages and a few of them are listed in Table 2 below. The flexibility and complexity ratings are subjectively assessed.



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Table 2, Overview of valvetrain system differences, numbers are approximate and may deviate for certain engine specifications

System Approximate

phasing range

Approximate

lift range

Flexibility Complexity Special

features

V-Tec Long dur. 260

Short dur.210

High 10 mm

Low 7 mm

Very limited,

only two or

three settings

Simple and

easily

controllable

Noticable

mode switch

Valvetroonic Possible

phasing

around 60

CAD relative

to crankshaft

0,3-9,9 mm Adjustable

intake and

exhaust valves,

large variation

span

Slightly

complex

installation,

advanced

control

Completely

unthrottled

operation

MultiAir Duration from

60 to 260 CAD

1-8 mm Only intake

valve valves

adjustable, large

variation span

Simple

installation,

advanced

control

Multiple

valve lifts

possible

Lotus AVT Limited by

collision

occurrences

Limited by

collision

occurrences

Practically

unlimited

Advanced

installation

and advanced

control

Possible to

specify

custom valve

lift curve

appearances

Cargine Limited by

collision

occurrences

3-12 mm Fully flexible,

but cannot

change valve

opening and

closing speed

Advanced

control

Time

constant

behavior

2.2 Literature survey

The literature survey is split into brief summaries of each paper with bullets describing the conclusions. The company, institute or university name is included, but for full reference see reference list.

1. Fault Diagnosis of Fully Variable Valve Actuators on a Four Cylinder Camless Engine, Robert Bosch GmbH [10]:

The paper focuses on FVVT diagnosis possibilities by using air intake system sensors such as Manifold Absolute Pressure, MAP, sensors. Results obtained on a 4-cylinder test bench engine are presented for the



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early intake opening strategy under different loads, and at medium range rotational speeds on steady-state conditions. Fault cases and the related emission impact are also described.

• Detection and identification of the different critical faults on each actuator is possible by using a Fourier series signal model of the MAP sensor

• Hydro Carbon emissions, HC, will increase 150% if one inlet valve or exhaust valve is malfunctioning and if no counteraction is applied

• Difference/drift in lift heights are typically a minor problem

• Inlet valve opening difference between cylinders upwards of 30°CA have only a slight impact on emissions, at 50°CA difference the effect is however substantial

• Exhaust valve closing difference between cylinders of more than 40°CA have a substantial impact on emissions

2. Unthrottled Engine Operation with Variable Intake Valve Lift, Duration, and Timing, General Motors [11]

The paper describes single cylinder testing performed at General Motors Research and Development, using a 3,4 liter double overhead camshaft, DOHC, engine that has been equipped with various camshafts with different lift and duration.



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Figure 10, Graphs from “Unthrottled Engine Operation with Variable Intake Valve Lift, Duration, and Timing” paper

• A 7 % net-specific fuel consumption improvement is reported for unthrottled engine operation using an early intake valve closing, EIVC, strategy

• For unthrottled engine operation, the fuel consumption improvement diminishes quickly by increasing the intake-valve duration, reducing the peak intake-valve lift, and/or retarding the intake-valve opening position



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• For the fuel-consumption-optimized unthrottled engine operation, the engine-out HC emission levels are 25 % higher and the engine-out Nitrous Oxide, NOx, emission levels are 25 % lower compared to the throttled engine operation with conventional and fixed valve timing

• Higher intake-charge velocities and in-cylinder swirl and tumble levels can be achieved with EIVC valve events; however, these enhanced charge motions require significant pumping work and eliminate the fuel consumption improvement of unthrottled engine operation

3. Unthrottled Engine Operation using Variable Valve Actuation: The Impact on the Flow Field, Mixing and Combustion, Loughborough University [12]

The paper describes single cylinder tests performed on a gasoline direct injection, GDI, engine using the Lotus AVT system and include comparisons with standard camshafts, EIVC strategies (Miller) and early closing of one of the inlet valves (Miller + Swirl). Comparisons in fuel consumption and emissions as well as optical surveys of the in-cylinder flow are also included as are detailed investigations related to start of injection, SOI, of the fuel.

• A significant reduction in indicated specific fuel consumption, ISFC, is exhibited when either both or only one inlet valve are operated under unthrottled operation

• The timing of the SOI in relation to the valve closure point plays an important role for engine performance

• SOI between 450° and 480° produced notably poor performance for all the unthrottled EIVC cycle profiles under consideration due to the sudden change in the air flow structures across the nozzle tip as the short duration cam profiles closed

• Although single valve profiles were capable of generating a strong swirl motion whilst open, this swirl ratio within the cylinder decayed rapidly once the valve had shut

• The work shows that fuel economy benefits can be gained by operating the engine with unthrottled EIVC operation. However, the interaction between the intake air and direct injection fuel spray means performance is highly dependent upon which valve is operated and the timing of the direct injection fuel spray



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4. Comparison between Unthrottled, Single and Two-valve Induction Strategies Utilizing Direct Gasoline Injection: Emissions, Heat-release and Fuel Consumption Analysis, University College London [13]

The paper describes single cylinder tests with fixed camshafts all with short duration and with varying lift heights. A standard production camshaft with a 9.1 mm lift was also compared to those with shorter durations and lower lifts. Both port fuel injected, PFI, as well as GDI injection systems were tested. Measurements of specific fuel consumption and exhaust emissions were carried out, while the combustion was analyzed using heat release analysis.

Table 3, Summary of results from “Comparison between Unthrottled, Single and Two-valve Induction Strategies Utilising Direct Gasoline Injection: Emissions, Heat-release and Fuel Consumption Analysis” paper

• The results show that there are worthwhile fuel consumption and exhaust emission benefits to be gained through de-activation of one of the two inlet valves at part-load conditions

• Single valve operation gave fuel consumption benefits of up to 5% at low load and 2.5% at medium load operation at the two engine speeds tested over the baseline case (9.1 mm camshaft with two-inlet valve operation)

• Generally all cases investigated with one-intake valve operation, showed faster heat release rates than for the equivalent two-inlet valve operation

5. Development of a Two-Stroke/Four-Stroke Switching Gasoline Engine – The 2/4SIGHT Concept, Ricardo UK Ltd [14]



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The paper is a concept study of a switching two-/four stroke engine. Different FVVT systems are also evaluated. The paper identifies the optimum combination of two-stroke and four-stroke operating regimes for fuel consumption, emissions and drivability as well as investigating potential mechanisms for switching between these combustion regimes in a gasoline engine for passenger car use. The potential benefits in fuel consumption, emissions and drivability are also quantified.

Figure 11, Graph from the paper Development of a Two-Stroke/Four-Stroke Switching Gasoline Engine – The 2/4SIGHT Concept

• The resulting concept engine is a 1.04 l in-line three-cylinder engine with an expected performance of 230 Nm and 85 kW

• When simulated in the Ford Focus, a European C-class car, the powertrain recorded a fuel consumption of 5.93 l/100km over the NEDC drive-cycle, without the need to use lean combustion

• The improvement in fuel efficiency is 24% compared with the series production Focus vehicle equipped with an equivalent conventional gasoline engine



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3 DESIGN AND IMPLEMENTATION

3.1 Engine specification

The cylinder head type used both as reference and with FVVT actuators mounted is based on the latest SAAB Ecotec engine. The engine has in its original state continuously variable cam phasing, wall guided direct injection, turbocharging as well as E85 flexfuel capability as standard. The high performance version of this engine has a maximum torque of 350 Nm and a maximum power of 260 hp while maintaining Euro 6 emission levels.

In this project the standard cylinder head has been cut off and modified to fit the single cylinder dyno. The engine is also built to have a 0,55 liter (per cylinder) displacement rather than the original 0,50.

In order to not be limited by ignition quality, the ignition coil was updated to a high energy version of 60 mJ. 2 spark plugs were used, one for low load (Denso SIP plug with heat range equivalent to 6 in the NGK scale) and one for high load (conventional spark plug from NGK with heat range 7).

The standard camshafts that were used for reference tests have the following data:

• Inlet valve duration of 245°CA

• Exhaust valve duration of 242°CA

• 60°CA cam phasing range (from pin position)

• MOP Pin position (default position) for the inlet valve and exhaust valve of 126°CA and 125°CA respectively

• Maximum valve lift of 10 mm

3.2 FVVT installation

The installation of the Cargine actuators has been simplified as much as possible. Basically the standard camshaft mountings have been milled down and an aluminum console has been adapted to facilitate correct actuator position in relation to valve stems. Brushless position sensors have been employed in order to measure valve lift. Pressurized air and oil is fed from mounting blocks above the actuators.

The installation can be seen in Figure 12 and Figure 13.



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Figure 12, Cylinder head installation

Figure 13, Valve mounting console assembly

For the updated actuators a different design with a valve cover was used, however the basic concept was kept.

4 EXPERIMENTAL SETUP

4.1 Instrumentation

The test setup used in the project is comprised of the following equipment:

• Dynamometer: The torque is measured by a strain gauge type load cell mounted on a console attached to the dyno. Basically the console is a lever arm with a known length and since the dyno is freely hinged on bearings all the force that is necessary to keep the dyno from spinning is



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transferred by the load cell. The load cell is calibrated using weights which have traceability towards national standards.

• Cylinder pressure measurement: For cylinder pressure measurement the uncooled 7 mm AVL GU21D is used. The sensor is mounted flush with the roof of the combustion chamber and is positioned close to the cylinder liner to ensure that pressure oscillations due to knocking are detected accurately. The signal is processed and logged during 400 consecutive cycles using an AVL Indimaster 670. The sensor has had approved calibration status throughout the tests.

• Pressure: Pressure transducers, relative and absolute pressure transducers in various ranges as suited for application. All have been calibrated using equipment with traceability towards national standards.

• Temperature: Thermocouples, both of type K and resistive Pt100 probes. The sensor bodies are not subject for calibration; however the measurement modules have all been calibrated using equipment with traceability towards national standards.

• Analogue inputs: All the measurement modules used for these tests have been calibrated using equipment with traceability towards national standards.

• Flow:

o Fuel flow: Due to the high pressure fuel system the standard AVL fuel weigh have been circumvented and instead a Promass 80 coriolis flow meter has been used. The meter has been calibrated using equipment with traceability towards national standards.

o Engine air flow: An ABB Sensyflow air meter is installed in the supercharging unit for the single cylinder dyno and is measuring the air flow going into the engine combustion chamber. This meter is not subject for calibration.

o Valve actuator air flow: A Manger-Wittman air meter is installed in order to measure air consumption for the valve actuators. The meter has been calibrated using equipment with traceability towards national standards.

• Emissions:

o Exhaust gas composition: A Horiba MEXA 9100 exhaust gas analyzer has been used for the determination of the emission levels from the engine. Calibration has been performed on a daily basis using span gases. Lambda value has been calculated using the Spindt equation.



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o Engine soot: An AVL 415S smoke meter is used in order to determine soot levels in the exhaust gas flow. The meter is not subject to calibration other than those performed during each measurement by the meter itself.

In addition to the test cell specific sensors 4 inductive position sensors [3] were mounted in the actuator console to measure valve lift. The sensors detect the valve cup position, see Figure 14. As the piston design was altered for the updated valve actuator, the position sensor signal was not accurate for these tests. However, the general characteristic of the lift profile and also the duration of the valve lift can be determined using the inductive sensor. In addition to the inductive position sensors the updated actuators has also built-in indication of 1 and 9 mm’s valve lift as obtained from the switching sensor mentioned in section 1.3.1.4.

Figure 14, Actuator with position sensor (red)

4.2 Calculations and assumptions

4.2.1 Compressor power

Throughout the course of these tests there has been no engine-driven air compressor connected to the engine in order to supply air to the actuators. Instead the air has been supplied by the pressurized air supply in the test facilities which in turn means that the parasitic losses of the compressor must be calculated. The method of calculating the compressor power is to some extent concept based; the efficiency of the compressor, the possibility of not cooling the compressed gas etc. all influences the result dramatically. The



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calculation of the compressor power has been based on the following concept suggested by Cargine:

• Variable displacement piston compressor used for supplying air to the actuators, similar to that of a standard air conditioning, AC, system

• Uncooled air supply to the actuators, limited to 150°C due to oil degradation problems

• Reuse of the pressurized air from actuators, thereby lowering the compressor ratio without lowering the force needed to open the valves. This also lowers the temperature after the compressor so that no cooling of the air going to the actuators is required.

This concept is not realized physically in any way on the test rig, and moreover the concept would have to be adapted and refined in order to assess the final parasitic losses. Buffering tanks and similar features are also likely to be the result of such an adaptation as would calibration and drivability compromises.

However, using the concept outline above and the assumptions and formulae listed below a good estimation can be made.

The formulae used in the calculation of the compressor power are:

[

[F 1]

[F 2]

The above mentioned formulae are standard thermodynamic relations that will result in a correction of the volume flow according to the temperature that would have been the result of the compression. This correction of volume mass flow is necessary as the tests are performed using cooled air as supplied from the facility’s air compressor and the concept that is to be emulated is based on a non-cooled gas. The piston compressor is likely to have a very high isentropic efficiency at low piston speeds.

The final calculations in order to determine the power required for the compressor are shown below:

comp

inairaircorr

in

comp

inincomp

T

Tmm

P

PTTT

*

1

1

&& =

−

+=

−

γ

γ



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[F 3]

Formula [F 3 is calculated for both the compression and expansion of the compressor. Combined with the mass flow and a set efficiency the total power required to drive the compressor can be calculated.

The mechanical efficiency has yet to be determined; however as no reliable data for the compressor mechanical losses has been found the mechanical efficiency is set to 70%.

4.3 Fuels

In this investigation two fuels have been used in order to assess any synergies between fuel properties and FVVT as well as establish a baseline for two fuels with widespread use today.

4.3.1 E10

10% ethanol content in automotive fuels are to be implemented in Europe, in some countries this has already been done. The fuel used in these tests has the following specification:

• Batch 1

o RON 95,4 and MON 85,2

o AFR = 13,92

o Density = 747,4 kg/m3

o Lower Heating Value = 41,1234 MJ/kg

o C vol% = 83,2228

o H vol% = 13,1792

o O vol% = 3,5980

o Cx = 1,0000

o Hy = 1,8872

o Oz = 0,0325

• Batch 2


o AFR = 13,91

−

−=

−− 12

11

1112

11

1

1κκκ VV

VpW



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o C vol% = 83,0504

o H vol% = 13,2395

o O vol% = 3,7102

o Cx = 1,0000

o Hy = 1,8997

o Oz = 0,0335

• Batch 3 (only used for supplementary study)


o AFR = 14,04



o C vol% = 84,90

o H vol% = 12,81

o O vol% = 2,29

o Cx = 1,0000

o Hy = 1,7983

o Oz = 0,0202

The fuel parameters used for calculations have been updated for each batch. The fuel quality from a combustion properties point of view has been considered to be equivalent between batch 1 and 2, while a correction for ISFC comparisons equivalent to the change in lower heating value has been used for batch 3. Additionally, fuel from batch 3 has only been used for part load conditions which in turn remove any effects of the higher knock resistance.

4.3.2 E85

E85 have been in use for some time and is a well-known renewable fuel. The fuel used in these tests conforms to the CEC RF E85 +20ºC Euro V standard for certification fuel.

• Batch 1

o AFR = 9,74




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o C mass% = 56,9941

o H mass% = 13,0774

o O mass% = 29,9285

o Cx = 1,0000

o Hy = 2,7343

o Oz = 0,3942

• Batch 2

o AFR = 9,78


o C mass% = 57,2368

o H mass% = 13,0706

o O mass% = 29,6926

o Cx = 1,0000

o Hy = 2,7213

o Oz = 0,3894

The fuel parameters used for calculations have been updated for each batch, but the fuel quality from a combustion properties point of view has been considered to be equivalent between batches.

Combining E85 and FVVT is an interesting concept due to the following reasons:

• Due to its lower stoichiometric AFR, E85 will not be as sensitive to variations in trapping efficiency due to valve timing fluctuations as E10 or regular gasoline

• If E85 is considered to be the main fuel, the geometric compression ratio can be substantially increased; yet with the FVVT the effective compression ratio may be lowered to enable good performance levels also when running gasoline

4.4 Test procedure

4.4.1 General

The difference between a single cylinder engine and a multi ditto is not overly large when it comes to the combustion properties, yet it is not trivial to extrapolate the findings from tests in the single cylinder environment to that of a full engine. The arguably most significant difference is that the boundary



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conditions are affected by the external gas exchange system; the multi-cylinder pulsating behavior in the exhaust and intake manifolds is not replicated in the single cylinder environment. Moreover, the friction losses and cooling of the cylinder and cylinder head are not fully duplicated. Hence, it is important to define suitable compromises that still capture the characteristics of that which is currently investigated. These compromises and simplifications are described in this section.

The resulting operating point settings are the product of extensive surveys performed. The surveys have not been conducted by means of a design of experiments, DoE, approach but instead by using experience and previous data.

Typically, when using short inlet valve durations, such as the Miller cycle, the point at which the valve should be closing coincides with the maximum speed of the piston. An example of this is shown in Figure 15.

-360 -300 -240 -180 -120 -60 0 60 120 180 240 300Crank Angle [deg]

Me

an_E

1 [m

m]

-2

-1

0

1

2

3

4

5

6

7

8

PC

YL

_M

ean

[bar

]

-3

0

3

6

9

12

15

18

21

24

27

Figure 15, Miller cycle and the piston speed (grey)

This unfortunate circumstance means that any variations in the inlet valve closing timing leads to severe fluctuations in lambda value as the trapping efficiency is varying. In turn this means poor stability and higher fuel consumption. This particular relation has led to further investigations in order to mitigate inlet valve fluctuations effect on trapping efficiency. This will be further commented under the results section.

Additional relevant remarks are:

• For all operating points the soot level is monitored, and set to a maximum of 0,5 filter smoke number, FSN



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• Lambda is kept at 1,00 for all operating points (stoichiometric operation)

• All reference tests performed using standard camshafts have been performed using camshaft settings from the multi-cylinder cam strategies of the original engine

4.4.2 Part load

The operating points for part load conditions are derived from the NEDC assuming a 1,6 liter engine and a vehicle curb weight of 1600 kgs. The points have been selected on the basis of which part of the cycle most fuel is used. As the single cylinder engine does not exhibit the same friction as a standard engine, indicated mean effective pressure, IMEP, is used instead of BMEP or torque and this also means that to some extent the engine displacement is not critical.

Part load points comprise the following loads and speeds:

• 1300 rpm and 4,12 bar IMEP




In part load conditions the stability of the engine has a direct correlation with the fuel consumption; this means that if the engine is running unstably the fuel consumption will also increase. In turn, this means that contrary to high load operating points the instability level is not a fixed limit but rather something that must be avoided as much as possible. Exhaust backpressure is controlled by a fixed valve setting in the exhaust system.

4.4.3 Full load

Full load points comprise the following loads and speeds:

• 1500 rpm full load naturally aspirated (103,0 kPa MAP)

• 1500 rpm full load supercharged

• 2000 rpm full load naturally aspirated (103,0 kPa MAP)

• 2000 rpm full load supercharged

The operating limits for the full load points are strictly defined, yet the comparison between different valve systems must be done carefully as it may not be relevant to focus only on the maximum performance. Different strategies may have different viability when a complete engine is considered, yet with this in mind the data can be interpreted and conclusions may be drawn from the tests. Also, with the supercharging handled by an external screw compressor,



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the resulting boost pressure is not determined by the available exhaust gas enthalpy but instead virtually infinite amounts of boost may be delivered to the engine resulting in unrealistic operating points. The operating limits for full load are listed here:

• Maximum supercharged MAP may never exceed the reference operating point by more than 10 kPa (realistic operating conditions)

• Maximum cylinder pressure is set to 120 bar (pmax + 2σ)

• The standard deviation of IMEP must be kept below 0,40 bar (combustion stability)

• Inlet manifold air temperature is kept at 25°C

• Exhaust backpressure is adjusted so that exhaust pressure is half of the boost pressure (at low engine speeds with turbocharger, an engine is typically run with positive PMEP)

• Maximum CO-emission level is set to 2% (in-cylinder lambda when running large overlap)

4.4.4 Idle

The idling test points are conducted in the same manner as the part load points described in section 4.4.2, however the speed and load is set to:



No fixed limit on combustion stability has been enforced for the idling operating points, but rather the achieved stability is the output from the tests.

Lower speeds may have been of interest, however the single cylinder engine is suffering from excessive engine rotational speed fluctuations at too low engine speeds even though the engine is equipped with a heavy flywheel.

4.4.5 Parasitic loss investigation

In order to compare the parasitic losses associated with the camshafts and the calculated losses for the pneumatic FVVT (see section 4.2) tests have been performed at different loads and engine speeds both with belt driven camshafts and pneumatically driven FVVT:

• 1000 – 3000 rpm, 2 bar IMEP

• 1000 – 3000 rpm, 8 bar IMEP

The camshafts have been driven by a toothed belt with a single mechanical tensioner and the camshaft settings have been:



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• For the 2 bar IMEP operating points: 125°CA inlet max open position, IMOP, and 126°CA exhaust maximum open position, EMOP, (PIN-position of the standard engine)

• For the 8 bar IMEP operating points: 100°CA IMOP and 106°CA EMOP

Indicated load has been the basis of the comparison, as the losses for the FVVT is calculated in the post-processing of the data and cannot be added to the brake torque in real-time. Not only is the load kept at the desired level between the camshafts and FVVT, but the FVVT valve timings have been adjusted so that the cylinder pressure trace is comparable with that from the tests performed with camshafts. This is done in order to make sure that the frictional losses derived from between the piston rings and engine cylinder liner, which are dependent on cylinder pressure, are kept as similar as possible between the tests. For higher load tests this methodology could not be used, as the steep valve opening and closing behavior of the FVVT resulted in dramatically different cylinder pressure trace characteristics.

There are infinite possibilities to vary the feed pressure to the actuators in order to minimize the air consumption and hence the compressor power requirement. In these tests the pressure levels have been determined by observing the stability of the valve events, i.e. to lower pressure to the point of valve behavior instability or the effects on the cylinder pressure trace due to lower valve lift make the comparison with camshafts invalid. The exhaust valve actuator pressure feed is defining the pressure ratio at which the compressor for valve actuator air supply must operate and the lower pressure feed for the inlet valve actuators only lowers the compressor power by the reduction of air mass flow that is the result of running the actuators at a reduced pressure.

4.4.6 Remarks to test procedure

There are several additional effects that will come into play when the concept typically tried out in a single cylinder environment is transferred to a serial project. Some of these concerns and areas of further investigations are listed below:

• While the compression ratio can be adjusted quite freely in the single cylinder test environment an actual engine that would need to be able to handle high ambient temperatures, different altitudes as well as variations in fuel quality. The compromises involved in the selection of engine compression ratio is therefore based on more than knocking behavior experienced in the context of this project

• The camshaft and the cam timings are based on the four cylinder turbocharged engine currently in production. This means that strategies that take transient behavior for the CVCP mechanisms under



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consideration are implemented. The camshaft timings are not in any way optimized for the running conditions experienced in the single cylinder dyno. Arguably the opposite may be stated for the FVVT, however the transient response for the FVVT is likely to be as instantaneous as a change in spark timing for example. This means that this particular consideration may not be as critical for the FVVT

• The instantaneous exhaust backpressure is not replicated in the single cylinder dyno. The foremost characteristic that is lacking is the absence of the blow down pulses normally generated by the other cylinders in a multicylinder engine, which on a four cylinder engine would mean a detrimental exhaust backpressure increase during the valve overlap

5 EXPERIMENTS

5.1 Part load operation

5.1.1 Summary

One of the main drivers for FVVT is the improvement in part load operation efficiency. Running without any throttling losses is a big step in improving the SI-engine efficiency that leaves the combustion itself as the sole area for further improvement. Throughout the course of this investigation, and specifically in the area of part load operation, the interaction between the FVVT valve lift events and the combustion properties has proved to be an area that definitely needs further investigation. A few of the part load operating points does not show a benefit of the FVVT even though the pumping loss reduction is clearly evident in the cylinder pressure trace –especially at lower engine loads. Instead unfavorable emissions and exhaust temperatures indicate that the combustion itself is not ideal. There are various plausible explanations to this, most of which are discussed in section 7. What can be achieved with the fully flexible valve train in terms of reduction of pumping losses and how this particular valve system interacts with the particular combustion system that has been used for this investigation needs to be addressed separately. Additional studies at the end of the project resulted in vast reductions in fuel consumption using very high internal EGR levels.

Mainly two strategies have been used in order to adjust the engine air mass flow without throttling; short inlet valve duration (Miller) and long inlet valve duration (Atkinson). These strategies are shown in Figure 16 and Figure 17. Typical valve lift curves for camshafts are also included as reference.



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0

5

10

15

20

25

30

0

5

10

15

20

25

30

35

-360 -270 -180 -90 0 90 180 270 360

Va

lve

Lif

t [m

m]

Cy

lin

de

r P

ress

ure

[b

ar]

Crank Angle Degrees [°CA]

Camshaft Pcyl FVVT Pcyl Inlet camshaft Exhaust camshaft

FVVT Inlet 1 FVVT Inlet 2 FVVT Exhaust 1 FVVT Exhaust 2

Figure 16, Typical 3 valve short inlet valve (Miller) strategy at 4,12 bar IMEP and 1300 rpm

0

5

10

15

20

25

30

0

5

10

15

20

25

30

35

-360 -270 -180 -90 0 90 180 270 360

Va

lve

Lif

t [m

m]

Cy

lin

de

r P

ress

ure

[b

ar]




Figure 17, Typical long inlet valve (Atkinson) strategy at 4,12 bar IMEP and 1300 rpm (note that the difference in exhaust backpressure is due to 0-level cylinder pressure difficulties related to the long inlet valve event)



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The high internal EGR operating point valve strategy is exemplified in Figure 18.

0

5

10

15

20

25

30

0

5

10

15

20

25

30

35

40

-360 -270 -180 -90 0 90 180 270 360

Va

lve

Lif

t [m

m]

Cy

lin

de

r P

ress

ure

[b

ar]




Figure 18, High internal EGR operating point using short inlet valve duration. Operating point at 1500 rpm and 4 bar IMEP

This strategy, due to being developed at the latter part of the project, has not been implemented for the operating points using E85 fuel.

Included in the evaluation is also a comparison between a FVVT 13,1 compression ratio engine and a camshaft based 9,2 compression ratio engine. As will be showed in section 5.2 the possibility of increasing compression ratio while maintaining performance of the engine is clearly one of the potentials for a FVVT engine, hence the comparison.

Summarized below are the best operating points for the FVVT, see Table 4 and Table 5.



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Table 4, FVVT vs camshafts part load consumption comparison

Table 5, FVVT CR 13,1 vs camshaft CR 9,2 part load consumption comparison

*) Fuel from batch 3 used, but ISFC comparisons corrected according to LHV

It should be noted that the data presented in this section are not corrected with any calculated parasitic loss from the auxiliary components necessary to power the FVVT nor are the friction losses from the camshaft drive included in the reference. Preliminary calculations regarding these losses can however be found in section 5.4.1.

It should also be noted that the comparative figures are based on a reference using CVCP and not fixed cam shafts. The reduction in fuel consumption when going from fixed camshafts to CVCP for these part load operating points is estimated to be in the range of 4-6%.

The engine efficiency and fuel consumption is based on the indicated work as derived from the cylinder pressure trace and is for that reason slightly higher than that calculated from the braked torque. It may also be interesting to know that the parasitic losses at the speeds used for the part load tests are similar between FVVT and camshafts, as seen in section 5.4.1.



Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


Page 44 (70)

5.1.2 Detailed data

The tests run at 1300 rpm and 4,12 bar IMEP are indicating a substantial increase in engine efficiency, see Table 6.

Table 6, 1300 rpm and 4,12 bar IMEP results

Not surprisingly, at the lower load of 1,37 bar IMEP at 1600 rpm the higher degree of throttling when running with camshafts is showing an even larger benefit for the FVVT. See Table 7.




Per-Erik Nilsson

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As previously mentioned, the internal EGR strategy was not used for the operating points with E85 fuel.

At the higher speed and load of 4,58 bar IMEP at 2250 rpm the situation is different; E10 is clearly more beneficial and in particular the longer inlet valve duration is the better strategy. At this operating point the use of internal EGR did not prove to be beneficial. See Table 8.


The instability of the tested reference point when running E10 is unexpectedly high, around 8% IMEP CoV, as can be seen in Table 12. This is mainly due to the extreme overlap of the reference camshaft settings which in combination with the exhaust system used in the single cylinder leads to knocking and excessive residual gasses in the cylinder. Obviously this camshaft strategy was never intended for the high compression ratio with this fuel, and so the spark timing is significantly retarded. The reason for why this camshaft strategy is used for the production engine is a compromise for better transient behaviour.

The operating point run with most setup and strategy variations is the 1500 rpm and 4,00 bar IMEP. See Table 9.



Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


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During the valve adjustment sweep performed with different internal EGR levels, the combustion seamlessly switched to HCCI mode. As this project did not include the scope of possible HCCI operation, while being a potential benefit of the FVVT, the boundary limits and controllability of this combustion mode was not further investigated

The results show that

• Even though there are some flow losses for the long inlet valve duration strategy, the strategy is an viable alternative to that of short inlet valve durations

• Advanced strategies that results in internal EGR bulking even though manifold pressure is at atmospheric conditions has proved to be



Per-Erik Nilsson

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Öyvind Gundersen

Date – Rev 20111130-4


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beneficial for short inlet valve durations, however fairly drastic valve timings have to be employed to do so

• With short inlet valve durations on lower speed and loads, using only one inlet valve is more beneficial than using two

• When using long valve durations both inlet valves should be used, regardless of load and speed

• The effects of different fuels are not fully consistent, additionally the internal EGR valve strategy has not been tested using E85 fuel

• The most significant fuel consumption reduction when considering this FVVT technology applied to this specific combustion system is obviously that of increased compression ratio enabled by the extended full load performance of the FVVT

In Table 10, Table 11, Table 12and Table 13 more detailed results from the part load operating points are summarized.

Table 10, 1300 rpm and 4,12 bar IMEP part load operating points




Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


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The asterisk (*) is also in this case highlighting the use of fuel from batch 3.

5.1.3 Tests performed with port fuel injection, PFI

In order to evaluate the potential difficulties with combining direct injection and FVVT, especially at extreme Miller cycling with significantly altered in-cylinder gas movement, some additional minor tests at 1500 rpm and 4,00 bar IMEP using port fuel injection were also carried out. In short, the results were:

• No clear improvement in fuel consumption with PFI versus DI at same operating conditions (i.e. same valve lift strategies)



Per-Erik Nilsson

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Date – Rev 20111130-4


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• Minor improvements in emissions, typically CO and O2 were reduced when running PFI due to enhanced fuel mixture preparation

• No change in soot levels

It should be pointed out that port fuel injection was never intended for this cylinder head and also that fuel mixture preparation, when using port fuel injection, is to some extent dependent on valve lift speed as the near-critical gas speed over the valve at low lifts effectively increases the blending of fuel and air.

5.2 Full load operation

5.2.1 Summary

Given all the degree of freedom available with the FVVT and the inherent difference in valve lift characteristic between FVVT and regular camshafts, any straight comparison between the two is difficult to perform. The most obvious is that of maximum performance achievable with the different valve systems and that is what is presented here.

One of the compromises involved is that of selection of valve spring stiffness. It is clear that the desire to run without throttling at lower engine speeds and loads requires stiffer valve springs in both inlet and outlet side. Inlet side need stiffer springs due to controllability, but in the case of exhaust valves the stiffer springs are typically needed in order to maintain closed valves when Miller cycle is used. The use of stiffer springs in conjunction with the high cylinder pressure at full load conditions results in the need of high feed pressure in order to open the valves sufficiently. Pressures upwards of 10-13 bar are typically needed in order to open the valves at the desired time with the desired lift height at the high load operating points presented in this report. The increase in pressure will also have a negative effect on the parasitic loss for the FVVT operation.

Due to the Atkinson cycle having a longer effective expansion stroke than compression stroke, it can potentially yield an increase in efficiency as well as enable a simplified variable compression ratio. This particular strategy did however prove to be unrealistic under the running conditions for this engine; the required boost pressure necessary to maintain engine output are too excessive.

5.2.2 Naturally aspirated

One of the most significant benefits of the FVVT is that the duration of the valve lift events can be altered. With CVCP a desired valve overlap may be achieved, but this would always affect both exhaust opening and inlet closing which may end up in an unfavorable gas exchange process. This is one of the main contributors to a lower pumping loss for FVVT overall in the full load operating points, both naturally aspirated and supercharged. Typically this would in turn



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increase the engine output equivalent to the reduction in pumping losses, which have proven to be in the range of 0,05 to 0,15 bar in mean effective pressure. However, the results show that while this pumping loss reduction is achieved, the combustion itself does not match the reference case with camshafts. Even though the traditional engine parameters such as start of injection, fuel pressure, spark advance etc. as well as the FVVT valve timings themselves have been optimized for the operating points this potential increase of performance cannot be realized using this particular combustion system. In Figure 19 this unforeseen behavior is clearly visible.

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Figure 19, 1500 rpm full load naturally aspirated with E10 fuel and 13,1 compression ratio

Again, the optimized exhaust valve opening, which in this case resulted in a reduction of pumping losses by 0,07 bar, does not fully compensate for the lower combustion pressure. The rate of combustion is clearly higher for the reference camshaft case, which is surprising giving the fact that the regular camshaft settings with more overlap than the FVVT settings would suggest an increase of residual gases and hence a slower combustion rate. The heat release rate is shown in Figure 20 which reaffirms the conclusion that the main difference between the operating points is the combustion rate.



Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


Page 51 (70)

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Figure 20, Heat release rate for 1500 rpm naturally aspirated full load

This behaviour is also found at 2000 rpm, as seen in Figure 21. For a summary of the performance and some additional parameters see Table 14.



Per-Erik Nilsson

Joakim Karlsson

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Date – Rev 20111130-4


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Figure 21, 2000 rpm full load naturally aspirated with E10 fuel and 13,1 compression ratio

Table 14, Detailed data for selected naturally aspirated operating points

Valve

system

Speed

[rpm] Fuel

IMEP

[bar]

PMEP

[bar]

Lambda

[-]

CO

[%]

NOx

[ppm]

Exhaust

temp [°C]

SOI

[°BTDC]

Ind ETA

[%]

FVVT 1500 E10 10,48 -0,08 1,001 0,875 2743 664 265 34,33

Cams 1500 E10 10,87 -0,15 1,001 0,497 2677 670 260 35,55

FVVT 1500 E85 12,68 0,05 1,001 0,645 3290 577 330 41,43

Cams 1500 E85 12,70 -0,15 0,995 0,729 2822 584 300 41,20

FVVT 2000 E10 11,13 -0,17 0,996 0,639 2892 712 305 36,82

Cams 2000 E10 11,50 -0,24 0,999 0,499 2997 704 280 37,68

FVVT 2000 E85 13,03 -0,04 1,000 0,746 3383 616 350 42,31

Cams 2000 E85 12,76 -0,26 0,996 0,723 2972 619 300 42,14

There are differences between the behaviour when running E10 and E85 as seen in Table 14. E85 would appear to have a more favourable combustion; however the effect of an optimized exhaust valve opening has a larger impact on E85 due to the higher cylinder pressure achieved by optimized spark advance so the difference in IMEP between E10 and E85 can mostly be attributed to this fact rather than improvements in combustion behaviour.



Per-Erik Nilsson

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Öyvind Gundersen

Date – Rev 20111130-4


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5.2.3 Supercharged

The most significant increase in performance with the FVVT can be found in the supercharged operating points, in particular at 2000 rpm. The IMEP plotted against manifold pressure, which gives an indication of the level of performance that is achieved for a certain boost pressure, can be found in Figure 22 and Figure 24.

Figure 22, Supercharged full load points at 1500 rpm

For 1500 rpm the largest difference between FVVT and regular camshafts is found when running E85. This is due to irregular combustion, i.e. pre-ignition occurances, when running the camshafts. When running the FVVT the lack of such irregular combustion may be attributed to the significant increase in valve overlap and hence better scavenging. In addition to this, the boost level could be further increased to the maximum allowed, which in turn also leads to more favourable scavening and also a higher level of positive pumping work. The significantly more rapid exhaust valve opening for the FVVT is also resulting in a much improved relation between initial pumping loss and expansion ratio around BDC. See Figure 23.



Per-Erik Nilsson

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Page 54 (70)

0,86 0,87 0,88 0,89 0,90 0,91 0,92 0,93 0,94 0,95 0,96 0,97 0,98 0,99 1,00Volume

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Figure 23, Zoomed up cylinder pressure diagram showing pressure around BDC

In Figure 23 the FVVT has less expansion (i.e. ”loosing” the beneficial area ’A’) but has a significantly lower pumping loss (area ’B’ versus area ’C’, as seen when using the enclosed area for each pressure trace). This is clearly a benefit seen in all full load points, but this is even more clear at boosted conditions as the pressure levels increase.

When evaluating the performance for the camshaft valve system, it is important to have in mind that the CVCP has a certain response time, and that this operating point is typically one at which the best steady state valve timing strategy cannot be used fully but must be a compromise towards transient response behaviour. As previously noted, this compromise would not apply to the FVVT since the FVVT has the possibility of changing valve timing dramatically from one cycle to the next.

At 2000 rpm the benefits of FVVT emerge even more clearly. This is an operating point even more dependant on boosted performance and knocking behaviour. It is also an operating point at which the scavenging strategy using valve overlap and positive pressure differential between inlet and exhaust ports

A

C B



Per-Erik Nilsson

Joakim Karlsson

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Date – Rev 20111130-4


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is even more pronounced. The performance versus the boost pressure can be found in Figure 24.

Figure 24, Supercharged full load points at 2000 rpm

For the operating points using E85 there is only a slight increase in performance that can be related to more efficient gas exchange. A good example is shown in Figure 25 at which the operating point at 170 kPa MAP is shown in more detail.



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Date – Rev 20111130-4


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Figure 25, 2000 rpm 170 kPa supercharged with E85 fuel and 13,1 compression ratio

As seen in Figure 25, the combustion is very similar between the two cases. Valve timing differs to some extent, and contrary to what has been seen previously the FVVT exhaust valve opening is not timed at optimum position. The reason for this is that the actuator cannot overcome the cylinder pressure before this rather late timing. The supplied feed pressure for the exhaust actuators at this specific operating point is 10,5 bar absolute pressure, and even though more pressure could be supplied to the actuator, the stability of the actuator is in turn adversely affected by excessive pressure levels. This limitation has been addressed in the updated actuator.

At 2000 rpm and E10 fuel with a relatively high compression ratio of 13,1 the potential of FVVT is even clearer. At these operating points it is primarily the resistance to knock that is determining the performance, and the foremost reason for the FVVT to excel for these operating conditions is the possibility of minor adjustments that enables further supercharging and hence increased scavenging. In Figure 26 the maximum performance operating point for the camshaft setup is compared with the FVVT. The operating conditions enabled the FVVT to have a relative volumetric efficiency of 104% (i.e. 4% scavenging) while the camshafts came to a full stop at 97%.



Per-Erik Nilsson

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Öyvind Gundersen

Date – Rev 20111130-4


Page 57 (70)

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Figure 26, 2000 rpm supercharged with E10 fuel, 130 kPa MAP and 13,1 compression ratio

Similar to the data presented in Figure 25, there are issues regarding exhaust valve opening behaviour. The lower lift height of the exhaust valve number 1 is an effect of this. The lift heights at these low speed full load points are however not critical; various lift heights were tested on the inlet valves without much effect on performance and even the relatively low exhaust valve lift of 5 mm is sufficient. The valve lift will however become critical as speed increases. This behaviour have however been addressed in the updated actuator.

Some overview data can be found in Table 15. Note the increase in CO for the FVVT operating point with increased boost level. This is a result of different in-cylinder AFR even though overall lambda is at 1.

Table 15, Detailed data for selected supercharged operating points at 13,1 compression ratio

Valve

system

Speed

[rpm] Fuel

MAP

[kPa]

IMEP

[bar]

Lambda

[-]

CO

[%]

NOx

[ppm]

Exhaust

temp [°C]

SOI

[°BTDC]

Ind ETA

[%]

FVVT 2000 E10 129,6 15,13 0,996 1,276 2467 751 307 35,32

Cams 2000 E10 130,3 14,97 1,003 1,075 3274 739 260 35,02

FVVT 2000 E10 158,3 19,57 0,994 1,784 1189 739 307 34,13

FVVT 2000 E85 170,8 22,96 0,996 1,731 1043 637 350 36,87

Cams 2000 E85 170,1 22,83 0,997 1,363 3457 683 300 38,84



Per-Erik Nilsson

Joakim Karlsson

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Date – Rev 20111130-4


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5.3 Idling

5.3.1 General

Idling performance of an engine is typically oriented towards combustion stability and so called Noise, Vibration and Harshness, NVH, requirements. In order to lower the engine load to 1 bar IMEP the camshaft setup requires a high degree of throttling and in the case of FVVT the lift heights and durations are lowered to extremely small valve lift events.

Short inlet valve durations are definitely more suitable than long inlet valve durations due to the fact that the valve lift height can be further reduced and hence the sensitivity to inlet valve closing timing variations can be minimized.

A typical idling operating point is shown in Figure 27. In order to facilitate throttle less operation the inlet valve lift is less than 1 mm and the timing is located close to TDC while still avoiding residual gases.

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Figure 27, Idling with E10 and 13,1 compression ratio

Another more exotic strategy was also used; very late inlet valve opening events which in fact occur during what is normally the compression stroke. See Figure 28.



Per-Erik Nilsson

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Date – Rev 20111130-4


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FVVT Pcyl Inlet camshaft Exhaust camshaft FVVT Inlet 1

FVVT Inlet 2 FVVT Exhaust 1 FVVT Exhaust 2

Figure 28, Idling with very late inlet valve event

The very late inlet valve event is not primarily reducing pumping losses, but due to the late inlet valve event the large scale in-cylinder turbulence is generated very close to the spark event. This leads to a rather quick combustion and hence vastly improved stability.

This strategy may be an enabler for running the engine without throttle, yet again it should be pointed out that this is not an optimization towards fuel economy due to the pumping losses induced. Naturally there will also be difficulties with mode switching and controllability, but this goes to prove the versatility of the FVVT system. The cylinder pressure trace is rather unlike that of a typical cycle, see Figure 29.



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Date – Rev 20111130-4


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0,0 0,1 0,2 0,3 0,4 0,5 0,6 0,7 0,8 0,9 1,0Volume

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Figure 29, Cylinder pressure trace for late inlet valve idling

In Figure 29 the fairly large pumping loss is clearly visible, yet one can intuitively understand that since the cylinder pressure is close to vacuum when the valve lift occur, the rush of air that enters the cylinder will not settle by the time the spark is igniting the mixture some 15 crank angle degrees after valve closing. Interestingly, the SOI was found to be at optimum in the normal timing range; i.e. in the initial part of what is normally the intake stroke. It may be that fuel vaporization is indeed improved by the lower pressure.

A summary of the idling results can be found in Table 16.



Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


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Table 16, Idling results for different FVVT strategies and camshaft references

Valve

system

Cam

strategy Fuel

MAP

[kPa]

IMEP

[bar]

IMEP

CoV [%]

PMEP

[bar]

Lambda

[-]

ISFC

[g/kWh]

Ind ETA

[%]

FVVT Short 2IV E10 100,0 1,03 26,80 -0,56 1,005 417,2 20,78%

FVVT Short 1IV E10 100,1 1,01 32,09 -0,55 0,992 428,2 20,34%

Cams PIN E10 25,6 1,01 2,23 -0,79 0,998 432,5 20,34%

FVVT Short 1IV E10 100,1 1,19 19,43 -0,54 1,002 379,7 22,93%

Cams PIN E10 27,1 1,22 1,64 -0,77 0,996 396,7 22,08%

FVVT Short 1IV E85 100,2 0,97 22,09 -0,55 0,996 646,7 19,10%

FVVT Short 2IV E85 100,1 1,22 9,26 -0,53 1,013 557,9 22,44%

FVVT Short 1 IV E85 100,2 1,21 8,13 -0,51 1,007 559,9 22,04%

FVVT Late 1IV E85 100,1 0,99 5,65 -0,84 1,009 682,3 18,43%

FVVT Late 2IV E85 100,2 1,02 4,44 -0,75 0,992 664,2 18,79%

Cams PIN E85 26,9 1,00 1,95 -0,82 1,000 659,1 19,17%

Cams PIN E85 28,8 1,19 1,34 -0,81 0,999 595,3 21,22%

Additional runs with camshafts at Ɛ = 9,2:

Overall, the results are showing that for this combustion system combined with the Cargine FVVT a fuel reducing unthrottled valve strategy cannot be implemented and still meet the combustion stability targets. If an engine concept without throttle is the main target, it can however be achieved by the late inlet valve strategy at some cost in efficiency and fuel economy (at idling conditions).

5.4 Overall analysis

5.4.1 Power consumption

The parasitic loss investigation methodology is described in 4.4.5 and the calculation and assumptions are described in 4.2. The resulting friction loss is shown in Figure 30 in which camshaft frictional losses as well as the frictional losses measured with and without the addition of the compressor power loss are presented.



Per-Erik Nilsson

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Page 62 (70)

0,4

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[b

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Engine Speed [rpm]

Camshafts 2Bar IMEP

Camshafts 8Bar IMEP

FVVT 2 BarIMEP (compr)

FVVT 8 BarIMEP (compr)

Figure 30, Friction loss as a function of engine speed

The results show that the parasitic losses that can be expected with the pneumatic FVVT are equivalent to or lower than that with the camshafts at both 2 bar and 8 bar IMEP at lower speeds, while at the higher speeds measured the increase in parasitic losses are upwards of approximately 10%.

These tests have been performed on low engine speeds and care must be taken to not extrapolate these results to higher engine speeds since the required lift height and thereby actuator feed pressure will increase to levels currently unknown. The air mass flow required to operate the FVVT will also scale with the engine speed, and this is clearly showing in Figure 30. However, these results assume that all 4 valves are to be used. It is likely that at some operating conditions the number of valves used may be reduced to 3 or even 2 valves with overall beneficial total engine efficiency. The exact trade-offs that would be the result of an optimization of FVVT parasitic losses and pumping losses are yet to be established, however the operating points showing the highest fuel consumption reduction have typically been found to be using a strategy with 3 valves per cylinder in operation which indicates an even further reduction in parasitic losses of the engine.



Per-Erik Nilsson

Joakim Karlsson

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Date – Rev 20111130-4


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The tests have been performed using the updated valve actuator. The valve springs have been selected so that stable valve lift behaviour can be maintained even at the very low lift heights required for unthrottled idling using the Miller strategy. This setup reflects a suitable final valve spring compromise.

The resulting friction loss of approximately 0,2 bar FMEP for the camshafts compared to the FVVT data without the added loss from the pneumatic system correlates well with in-house references.

5.4.2 Reference engine versus high compression FVVT engine

Additionally it may be of interest to compare a standard engine running at a compression ratio of 9,2 with a FVVT engine running at a compression ratio of 13,1. It has been showed that the FVVT can, to some extent, facilitate an increase in compression ratio while still achieving good full load engine output. However, the high compression ratio inevitably leads to unfavourable spark advance and poor efficiency at full load operation. See Figure 31.

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Figure 31, 2000 rpm supercharged comparison between FVVT and camshafts at different compression ratio

The penalty at full load operation in terms of efficiency is around 1,0-1,5% and the benefits at part load conditions is around 2,0-3,0% depending on load and speed. Detailed data for further comparison can be found in Table 10, Table 11, Table 12 and Table 13 in the part load section of this report.



Per-Erik Nilsson

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Date – Rev 20111130-4


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5.4.3 Valve stability effects

As previously mentioned in the report, the inlet valve closing timing is of outmost importance when the Miller cycle is used. This is due to the piston speed and induction stroke being at its strongest at the same time that inlet valve is closed, see Figure 15. Obviously, since the pneumatically driven FVVT is mechanically disconnected from the engine crankshaft it is interesting to evaluate the fluctuations and correlate them with fluctuations in AFR and IMEP.

The controllability of the valves is imperative when using extreme Miller or Atkinson cycle. During the testing various techniques have been employed in order to manually adjust the valves so that a suitable valve lift behavior is achieved. Hysteresis and non-linearity of the valve timings may clearly be a challenge for the engine control system.

Typically the standard deviation in crank angle degrees for the closing of the inlet valves is between 1-2°CA depending on valve pressure feed. The standard deviation of the valve closing has however very little, if any, correlation with IMEP fluctuations however the variations in lambda value cannot be disregarded. The lambda value standard deviation for a FVVT operating point can be upwards of 5 times that of a reference camshaft ditto. This may to some extent affect the fuel consumption.

5.4.4 Direct injection versus port fuel injection and other combustion system related discussions

A question worth serious consideration is that of which injection system that would be most suitable for a combustion system using extreme Miller cycling. The in-cylinder turbulence and fuel preparation are both dramatically altered, and if single inlet valve operation is used, the 4-valve tumbling gas movement will be changed to a swirl with the spark plug in the middle of the combustion chamber roof. This may not be an ideal concept for part load conditions, but would of course have to be related to the advantage a centrally mounted spark plug yields at full load operating conditions that are limited by knock (the longer the combustion flame have to travel the longer the air fuel mixture in the periphery of the combustion chamber have to self-ignite, hence a centrally mounted spark plug minimizes this distance).

Overall, the design of the combustion system may need a new approach if Miller cycling is to be implemented. Additionally, the use of high levels of internal EGR should be further investigated.

5.4.5 Higher engine speed and FVVT

As has already been shown under section 1.3.1.1 and in Figure 6 the valve lift speed is obviously not constant in terms of mm per crank angle, but rather a function of the physical properties of the actuator, feed pressure, valve spring



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and valve themselves. In this project there has been no tests performed at engine speeds higher than 3500 rpm, and so the behavior of the valves at the typical maximum SI-engine speeds of 6000 rpm and beyond cannot be stated here. Initial tests indicated issues with valve lift behavior and pressure level requirements at higher engine speeds and loads. After the actuator upgrade the problems were solved and the stability was acceptable within the tested load and speed range. See Figure 32.

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FVVT Pcyl Inlet camshaft Exhaust camshaft FVVT Inlet 1

FVVT Inlet 2 FVVT Exhaust 1 FVVT Exhaust 2

Figure 32, 3500 rpm full load naturally aspirated

Further work and optimizations may be required at higher engine speeds as difficulties with regards to timing and pressure levels will be more evident. There are however at this point nothing indicating that high speed operation will not be feasible.

6 CONCLUSIONS Part load:

• The part load operating points show a significant benefit of the FVVT

• At 4,12 bar IMEP and 1300 rpm the results show a significant increase in engine efficiency (-4,26% ISFC ) for the FVVT compared to camshafts

• At 1,37 bar IMEP and 1600 rpm, the higher degree of throttling when running with camshafts is showing an even larger benefit (-10,71% ISFC) for the FVVT



Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


Page 66 (70)

• At 4,58 bar IMEP and 2250 rpm and using a more conventional Miller timing (i.e. not high internal EGR strategy) the FVVT is better than camshafts (-5,00% ISFC ); E10 is clearly more beneficial and in particular the longer inlet valve duration is the better strategy

• At 4,00 bar IMEP and 1500 rpm the FVVT is significantly better than camshafts (-4,39% ISFC )

• The improvement is achieved with a comparison with an engine using CVCP. A CVCP system typically reduce the fuel consumption in these part load operating points by 4-6% depending on load

• Even though the pumping loss reduction is clearly evident in the cylinder pressure trace for practically all part load operating points performed with the FVVT, the fuel consumption does not necessarily follow the same trend

• The use of longer inlet valve durations may be better for fuel preparation and in-cylinder turbulence, however the longer inlet valve strategy has not been proven to work well in combination with the high internal EGR strategy

• Advanced strategies that results in internal EGR bulking even though manifold pressure is at atmospheric conditions has proved to be very beneficial for short inlet valve durations

• With short inlet valve durations on lower speed and loads, using only one inlet valve is more beneficial than using two

• When using long valve durations both inlet valves should be used, regardless of load and speed

• The effects of different fuels are not fully consistent, however for most operating points E85 is likely the more suitable fuel. This could possibly be the result of less sensitivity to AFR fluctuations

• There are no clear differences in fuel consumption with PFI versus DI at same operating conditions (i.e. same valve lift strategies)

Full load:

• At full load the pumping losses can be reduced by 0,05 to 0,15 bar in mean effective pressure. However, the results show that while this pumping loss reduction is achieved, the combustion itself does not match the reference case with camshafts.

• The most significant increase in performance with the FVVT can be found in the supercharged operating points, in particular at 2000 rpm.



Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


Page 67 (70)

• At 2000 rpm, E10 fuel and a compression ratio of 13,1 the resistance to knock is determining the performance. FVVT enables further supercharging and hence increased scavenging; the relative volumetric efficiency can with FVVT be raised to 104% while the camshafts end up at 97%.

• Overall, the FVVT enable higher performance and efficiency for the same boost level

Idling:

• Extremely late inlet valve event strategy may be an enabler for running the engine without throttle, however the pumping losses induced will result in an efficiency and fuel consumption loss

Parasitic losses:

• The parasitic losses for the FVVT are comparable to the camshafts for both 2 bar and 8 bar IMEP at lower speeds, while at the higher speeds the increase in parasitic losses are upwards of 10%

Engine flexibility:

• It has been shown that the FVVT can facilitate an increase in compression ratio while still achieving good full load engine output. However, the high compression ratio inevitably leads to unfavourable spark advance and poor efficiency at full load operation

• The most significant fuel consumption reduction when considering this FVVT technology applied to this specific combustion system is that of increased compression ratio enabled by the extended full load performance of the FVVT

7 RECOMMENDATIONS AND OUTLOOK • The combustion system must be revised and designed with the FVVT

valve timing strategies in mind for the pumping loss reduction to end up as a fuel consumption improvement. Exactly which parameters to look into will need to be clarified, but it is very likely that differences in in-cylinder gas movement and the combustion efficiency should be the main focus. This could be realized by combining additional tests and computational fluid dynamics, CFD, tools. Experiences obtained using high levels of internal EGR should also be included here

• Additional surveys of the effects of internal EGR bulking should be conducted



Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


Page 68 (70)

• The improvement potential of advanced cold start emission strategies enabled by the FVVT needs to be investigated on a multi cylinder engine

• The potential of improved transient behavior due to the instantaneous response of the FVVT needs to be investigated on a multi cylinder engine

• The exact trade-offs that is the result of an optimization of FVVT parasitic losses versus pumping losses and combustion efficiency should be investigated on a multi cylinder engine

• The potential of skip cycle/cylinder deactivation needs to be investigated on a multi cylinder engine

• Further work on the control strategies for the valves, especially when running at extreme valve timings is necessary due to the hysteresis and non-linearity of the valve events

8 REFERENCES

8.1 Web

[1] http://www.acea.be/

[2] http://www.erec.org/

[3] http://www.microstrain.com/ncdvrt.aspx

8.2 Literature

[4] Sugiyama Hiyoshi et.al., “Technology for improving engine performance using variable mechanisms”, SAE paper 2007-01-1290, 2007

[5] Henrik Dembinski et.al. “Miller-cycle on a heavy duty diesel engine”, Master of Science Thesis MMK2 2009:1 MFM124, 2009

[6] J. Ma, H. Schock, U. Carlson, A. Höglund, & M. Hedman, “Analysis and Modeling of an Electronically Controlled Pneumatic Hydraulic Valve for an Automotive Engine”, SAE paper 2006-01-0042, 2006

[7] Installation instructions for Free Valve Technology, Cargine Engineering AB, Ängelholm, 2008

[8] Öyvind Gundersen, “Free Valve Technology”, Master of Science Thesis MMK 2009:75 MFM130

[9] Joakim Karlsson, “Control Strategy for Fully Flexible Valve Technology”, Master of Science Thesis MMK 2010:18 MFM135



Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


Page 69 (70)

[10] Ipec Sarac et.al., “Fault Diagnosis of Fully Variable Valve Actuators on a Four Cylinder Camless Engine”, SAE paper 2008-01-1353

[11] David Cleary et.al., “Unthrottled Engine Operation with Variable Intake Valve Lift, Duration, and Timing”, SAE paper 2007-01-1282

[12] Philip A. Stansfield et.al., “Unthrottled Engine Operation using Variable Valve Actuation: The Impact on the flow Field, Mixing and Combustion”, SAE paper 2007-01-1414

[13] Rishin Patel et.al., “Comparison between Unthrottled, Single and Two-valve Induction Strategies Utilising Direct Gasoline Injection: Emissions, Heat-release and Fuel Consumption Analysis”, SAE paper 2008-01-1626

[14] R.J. Osborne et.al., “Development of a Two-Stroke/Four-Stroke Switching Gasoline Engine – The 2/4SIGHT Concept”, SAE paper 2005-01-1137

9 NOMENCLATURE Air Condition, AC

The European Automobile Manufacturers Association, ACEA

Air Fuel Ratio, AFR

Actuator Piston, AP

Brake Mean Effective Pressure, BMEP

Crank Angle, CA

Computational Fluid Dynamics, CFD

Continuously Variable Cam Phasing, CVCP

Design of Experiments, DoE

Double Overhead Camshaft, DOHC

Exhaust Gas Recirculation, EGR

Early Intake Valve Closing, EIVC

Engine Management System, EMS

Filter Smoke Number, FSN

Fully Variable Valve Timing, FVVT

Gasoline Direct Injection, GDI (DI)

Non Methane Hydro Carbon (emissions), NMHC (HC)

Indicated Mean Effective Pressure, IMEP



Per-Erik Nilsson

Joakim Karlsson

Öyvind Gundersen

Date – Rev 20111130-4


Page 70 (70)

Inlet Max Open Position, IMOP

Inlet Valve Closing, IVC

Inlet Valve Opening, IVO

Indicated Specific Fuel Consumption, ISFC

Lift Solenoid, LS

Manifold Absolute Pressure, MAP

New European Driving Cycle, NEDC

Nitrous Oxide, NOx

Noise, Vibration and Harshness, NVH

Port Fuel Injected, PFI

Start Of Injection, SOI

Spark Ignited, SI

Timing Solenoid, TS

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