The Miller Cycle on Single-Cylinder and Serial Configurations of a Heavy-Duty Engine

VARUN VENKATARAMAN

Master of Science Thesis Stockholm, Sweden 2017

Varun Venkataraman

Master of Science Thesis TRITA-ITM-EX 2018:28 KTH Industrial Engineering and Management Machine Design, Division of Internal Combustion Engines SE-100 44 STOCKHOLM

Examensarbete TRITA-ITM-EX 2018:28

Millercykeln i en Encylindrig och Flercylindrig Lastbilsmotor

Varun Venkataraman Godkänt Examinator Handledare 2017-12-22 Anders Christiansen Anders Christiansen Erlandsson Erlandsson Uppdragsgivare Kontaktperson Anders Christiansen Erlandsson Sammanfattning I jämförelse med sina föregångare, har moderna lastbilsmotorer genomgått en betydande utveckling och har utvecklats till effektiva kraftmaskiner med låga utsläpp genom införandet av avancerade avgasbehandlingssystem. Trots att de framsteg som gjorts under utvecklingen av lastbilsmotorer har varit betydande, så framhäver de framtida förväntningarna vad gäller prestanda, bränsleförbrukning och emissioner behovet av snabba samt storskaliga förbättringar av dessa parametrar för att förbränningsmotorn ska fortsätta att vara konkurrenskraftig och hållbar. Utmaningen i att uppfylla dessa till synes enkla krav är den invecklade, ogynnsamma balansgång som måste göras mellan parametrarna.

Förbränningsmotorns kärna är förbränningsprocessen, som i sin tur är kopplad till motorns luftbehandlings- och bränsleregleringssystem. I denna studie undersöks Millercykeln som en potentiell lösning till att nå de motstridiga kraven för framtida lastbilsmotorer, framförallt med fokus på potentialen att förbättra prestandan samtidigt som NOx-emissionerna hålls på konstant nivå.

Traditionellt har utvärderingen av Millercykeln utförts på encylindriga forskningsmotorer, vilket också har utgjort utgångspunkten i denna studie. Även om studier på flercylindriga simuleringsmodeller och forskningsmotorer har gjorts med konstanta inställningar för Millercykeln, så utförs de inte i samband med undersökningar av encylindriga motorer. Dessutom så möts inte kraven från insugssystemet på samma sätt mellan de olika motorkonfigurationerna.

Denna studie undersöker och jämför potentialen för ökad prestanda med Miller-cykeln mellan encylindrig och flercylindrig motorkonfiguration för en lastbilsmotor med ett tvåstegs turboladdningssystem, som representerar ett realistiskt insugssystem som möjliggör implementeringen av Millercykeln. För att undersöka motorprestationen så används i denna studie den kommersiella mjukvaran GT-Power.

Ytterligare resultat från studien innefattar kvantifiering av prestandakraven för ett högeffektivt tvåstegs turboladdningssystem och dess inverkan på temperaturen i inloppet till avgasbehandlings-systemet. En kvalitativ förståelse av betydelsen av interaktionen mellan cylindrar och effekten på cylinder-cylinder variationer med Millercykel utfördes också i simuleringar med flercylindrig motorkonfiguration.

Studien utvärderade Millertiming inom ett intervall på -90 till +90 graders vev vinkel från utgångsvinkeln för stängning av insugsventilen. Utvärderingen utfördes vid systemjämvikt vid en fullastpunkt (1000RPM), där basfallet för både encylindrig och flercylindrig motor för utvärdering av Millercykeln var det välkända fallet med konstant specifik NOx. Ett ytterligare fall framhäver NOx-reduktionspotentialen med Miller vid konstant EGR-flöde på en encylindrig konfiguration. Fallen med ökad prestation realiserades genom att öka lufttillförseln, bränslemängden och det geometriska kompressionsförhållandet.

Maximal prestandaökning observerades i fallet med ökad bränslemängd, och endast i detta fall utvärderades även konfigurationen med fler cylindrar för jämförelse av prestationsförbättringen med en encylindrig motsvarighet med Millertiming.

Den flercylindriga motorn innefattade EGR som en lågtryckskrets, och medan detta antagande förenklade i avseende på modellering och kontroll, så var det till fördel för konfigurationen med en flercylindrig motor (jämfört med encylindrig) på grund av reducerade pumpförluster. Som påföljd gjordes en jämförande undersökning med encylinder-modellen med motsvarande mottryck för flercylinder-modellen inställt som gränsvärde. Resultaten visar att encylinder- modellen representerar medelvärdet för cylindrarna i flercylinder-motorn när lämpliga gränsvillkor tillämpas som kontrollparametrar. Studien ger en grund för jämförelse av Millertiming på encylindrig samt flercylindriga konfigurationer, samtidigt som kraven på insugssystemet fastställs och utgör en utgångspunkt för att utvärdera Millercykeln och bestämma insugssystemets krav för hela motorns arbetsområde.

Master of Science Thesis TRITA-ITM-EX 2018:28

The Miller Cycle on Single-Cylinder and Serial Configurations of a Heavy-Duty Engine

Varun Venkataraman Approved Examiner Supervisor 2017-12-22 Anders Christiansen Anders Christiansen Erlandsson Erlandsson Commissioner Contact person Anders Christiansen Erlandsson Abstract Modern heavy-duty engines have undergone considerable development over their predecessors and have evolved into efficient performance machines with a reducing emission footprint through the incorporation of advanced aftertreatment systems. Although, the progress achieved in heavy-duty engine development has been significant, the future expectation from heavy-duty engines in terms of performance, fuel consumption and emissions stresses the need for rapid large-scale improvements of these metrics to keep the combustion engine competitive and sustainable. The challenges in resolving these apparently straightforward demands are the intricate unfavourable trade-off that exists among the target metrics. The core of the combustion engine lies in the combustion process which is inherently linked to the air handling and fuel regulating systems of the engine. This study explores adopting the Miller cycle as a potential solution to the conflicting demands placed on future heavy-duty engines with an emphasis on the performance enhancement potential while keeping the specific NOX emission consistent. Traditionally, evaluation of the Miller cycle is performed on single-cylinder research engines and formed the starting point in this study. While studies on full-engine simulation models and test engines with fixed Miller timing have been evaluated, they appear to be performed in isolation of the favoured single-cylinder approach. Additionally, the charging system requirements are not consistently addressed between the two approaches. This study investigates and contrasts the performance enhancement potential of the Miller cycle on single-cylinder and serial engine models of a heavy-duty engine along with a two-stage turbocharging system to represent a realistic charging system that enables implementation of Miller timing. The commercial engine performance prediction tool GT-Power was used in this study. Additional outcomes of the study included quantifying the performance demands of a high efficiency two-stage turbocharging system and its impact on the inlet temperature of the exhaust aftertreatment system. A qualitative understanding of the significance of cylinder interaction effects on cylinder-cylinder variations with Miller timing was also performed on the serial engine cases.

The study evaluated Miller timing within a range of -90 to +90 CAD from the baseline intake valve close angle. The evaluation was performed at steady-state operation of the engine at one full load point (1000RPM) wherein both the single-cylinder and serial engine Miller evaluation included a base case which characterises the Miller effect for constant specific NOX. An additional case highlights the NOX reduction potential with Miller for a constant EGR rate on the single-cylinder configuration. The performance enhancement cases were realised by increasing the air mass, fuel mass and the geometric compression ratio. Maximum performance increase was observed in the increased fuel mass case and only this case was evaluated on the serial engine for contrasting single-cylinder and serial engine performance enhancement with Miller timing. The serial engine incorporated EGR as a low-pressure circuit and while this simplified modelling and controller considerations, it led to biasing of results in favour of the serial engine configuration (over the single-cylinder) due to reduced pumping loss. A subsequent comparison case was evaluated on the single-cylinder model with backpressure settings from the serial engine model. The results show that the single-cylinder model is representative of the cylinder averaged responses of the serial engine when appropriate boundary conditions are imposed as controller targets. The study provides a basis for contrasting Miller timing on single-cylinder and serial configurations while determining the charging system requirements and presents a starting point to evaluate Miller timing and determine air system demands over the entire engine operating range.

5 6 FOREWORD

I would like to start by thanking my parents, Mala and Ramesh for their constant love, support and belief in my endeavours through the years. Their emphasis on a sound education was the prime enabler for me to have performed my education at a prestigious school like KTH and to have lived in a country I have come to admire as much as my own. To my supervisor, Anders Christiansen Erlandsson, who not only provided me an immensely interesting and challenging thesis proposition but also provided multiple opportunities to contribute and learn from the educational and research activities at the Division of Internal Combustion Engines. I am grateful for his encouragement and guidance which helped me define this thesis and acquire a deeper understanding into the fascinating field of combustion engines. I would like to thank the staff and doctor students at the division, Nicola Giramondi, Ted Holmberg, Arun Karuppasamy, Bertrand Kerres, Senthil Krishnan, Tara Larsson and Sandhya Thantla for providing a wonderful work environment and helping me with the numerous queries, literature references and proofreading that emanated through my thesis work. I would also like to thank Johanna Olsson and Gunilla Hugosson for their support on administrative issues and providing me access to the fika and supplies room for over a year! Many thanks to Björn Finér, who provided me a great workstation and promptly resolved system and network issues. I must also thank my little sister, Boo and my extended family for their love and support. I am grateful to friends old and new, who have helped and accompanied me not only through this master program but through the journey of life so far. Although not directly related to my thesis work, I would also like to thank the KTH Opportunities Fund for their generous fee waiver through my final study term.

Varun Venkataraman

Stockholm, July 2017

7 8 NOMENCLATURE

Notations

Symbol Description

L/D Lift over Diameter

Abbreviations

AFR Air-Fuel Ratio AHRR Apparent Heat Release Rate BDC Bottom Dead Centre BMEP Break Mean Effective Pressure BSFC Break Specific Fuel Consumption BTE Break Thermal Efficiency CA Crank Angle CAC Charge Air Cooler CARB California Air Resources Board CAx Crank angle at x% burned fuel mass CO Carbon Monoxide CPOA Cylinder Pressure Only Analysis DCR Dynamic Compression Ratio DCRM Diffusion Combustion Rate Multiplie EAT Exhaust Aftertreatment EGR Exhaust Gas Recirculation EIVC Early Intake Valve Close EOI End of Injection ERM Entrainment Rate Multiplier EVO Exhaust Valve Open FMEP Friction Mean Effective Pressure GA Genetic Algorithm GCR Geometric Compression Ratio GT Gamma Technologies

9 HC Hydrocarbons HDD Heavy-Duty Diesel HP High Pressure HRR Heat Release Rate IDM Ignition Delay Multiplier IMEP Indicated Mean Effective Pre ISC Interstage Cooler ISNO Indicated Specific NO ITE Indicated Thermal Efficiency IVC Intake Valve Close LIVC Late Intake Valve Close LP Low Pressure MEP Mean Effective Pressure MIMO Multiple Input Multiple Output

BSNOX Break Specific NOX

NOX Oxides of Nitrogen NVH Noise Vibration and Harshness PCP Peak Cylinder Pressure PCRM Premixed Combustion Rate Multiplier PMEP Pumping Mean Effective Pressure RPM Revolutions Per Minute SCR Selective Catalytic Reduction SISO Single Input Single Output SOI Start of Injection T2S Two-Stage Turbocharger TDC Top Dead Centre TPR Total Pressure Ratio VGT Variable Geometry Turbine VVA Variable Valve Actuation VVT Variable Valve Timing

10 Contents Sammanfattning 1

Abstract 4

Foreword 7

Nomenclature 9

Table of Contents 11

Introduction 13 Two-Stage Turbocharging ...... 13 The Miller Cycle ...... 14 Two-Stage Turbocharging with Miller ...... 16

Problem Description 17

Research Questions 17

Methodology 17 Predictive Diesel Combustion Model ...... 18 Basis of Miller Evaluation ...... 18 Single-Cylinder Miller Evaluation ...... 18 Serial Engine-T2S Miller Evaluation ...... 19 Limitations ...... 19

Engine Modelling Approach and Considerations 19 Introduction to GT-Power ...... 20 Base Engine Description ...... 20 Base Engine Model Capabilities ...... 20 Predictive Combustion Model Calibration ...... 21 Predictive NOX Model Calibration ...... 26 Single-Cylinder Miller Evaluation ...... 27 Serial Engine Miller Evaluation with T2S ...... 29

Results 31 The Miller Cycle on a Single-Cylinder Engine Conﬁguration ...... 31 Miller Effect for Constant ISNO ...... 31 Miller Effect for Constant EGR Rate ...... 39 Performance Enhancement by Increasing Air Mass ...... 42 Performance Enhancement by Increasing Fuel Mass ...... 46 Performance Enhancement by Increasing Compression Ratio ...... 50 The Miller Cycle on a Serial Engine Conﬁguration with T2S ...... 50 Miller Effect for constant ISNO ...... 50 Miller and Cylinder-Cylinder Interaction ...... 56 Cylinder-Cylinder Variation: Serial Engine Base Case ...... 56 Performance Enhancement by Increasing Fuel Mass ...... 57 Cylinder-Cylinder Variation: Serial Engine Performance Case ...... 60

Conclusions 60

Discussion and Future Work 62 DI Pulse Calibration ...... 62 Base Engine Friction Characterisation ...... 63 Intake Valve Proﬁles for Miller Timing ...... 63

11 Controller Considerations ...... 63 T2S Modelling and EGR Circuit Considerations ...... 64 Cylinder Interaction Investigation ...... 64

References 65 Appendix A1

12 The Miller Cycle on Single-Cylinder and Serial Conﬁgurations of a Heavy-Duty Engine Varun Venkataraman Division of Internal Combustion Engines KTH-Royal Institute of Technology Stockholm, Sweden

Anders Christiansen Erlandsson Supervisor and Head of the Division of Internal Combustion Engines KTH-Royal Institute of Technology Stockholm, Sweden

INTRODUCTION cation of two-stage turbocharging, dual loop exhaust gas recirculation (EGR) and combustion optimization. The future demands for heavy-duty engines are typi- The fuel consumption benefit can be attributed to a cally along the lines of lower emissions and fuel con- 1.4% improvement due to downsizing and two-stage sumption while maintaining or increasing the perfor- turbocharging while optimizing the air system (two- mance. Although these goals may appear clear, the stage and EGR) lead to an additional 5.2% improve- means of achieving them could be regarded as an ment. In total, 6.6% of the 7.5% fuel consumption enormous challenge for heavy-duty engine develop- improvement at the operation point of interest was a ment given that they share an unfavourable trade-off. result of the combined effects of downsizing with opti- mized air handling (charging and EGR) while the per- Downsizing of engines has been considered a plausi- formance and emissions were maintained. ble approach and has been widely implemented in the passenger vehicle segment. As performance cannot The contrast between [1] and [3] shows that downsiz- be compromised, the air demands of the downsized ing with the combination of two-stage charging alone engine requires additional boost that is generally pro- would not suffice to justify the fuel consumption ben- vided by a two-stage charging system. This corre- efit. It is the thorough optimization of the air system spondingly leads to an increase in the the specific for the design point(s) that enables a greater benefit power output, the peak cylinder pressure (PCP) and by tapping into reduced pumping loss in addition to no major friction reduction benefit over the base en- lower friction. gine (medium-duty truck) unless the number of cylinders are reduced to achieve the lower displacement It appears that two-stage turbocharging potentially [1]. has a key role to play in addressing the demands of future heavy-duty engines and hence a brief overview Engine-vehicle simulation studies [2] on a heavy-duty of the concept is provided. vehicle have shown a 2-5% fuel consumption benefit under the CARB drive cycle when downsizing from six to five cylinders and using the baseline turbocharger Two-Stage Turbocharging Two-stage tur- (one-stage) with a turbocompound system. An eval- bocharging refers to the use of two turbochargers ar- uation of variable valve timing (VVT) on the baseline ranged in series wherein the exhaust gas passes over engine to enable Miller and Atkinson cycles showed two turbines, each of which powers a compressor to no more than a 1% fuel consumption benefit. VVT boost the intake air twice. The total pressure ratio appeared to be more promising for exhaust thermal (and expansion ratio) is a product of the individual management according to this study. stage values. The turbocharger operating with its turbine immediately after engine exhaust is referred to The favoured approach to downsizing is not unani- as the high pressure (HP) stage and the one that fol- mous as [3] shows that fuel consumption can be re- lows downstream is the low pressure (LP) stage. duced by 7.5% at a typical heavy-duty engine operating point (1000RPM and 35% load which corresponds The purpose of two-stage turbocharging is to achieve to 80 km/h on highway) through the combined appli- high boost pressures while maintaining reasonable

13 stage pressure and expansion ratios which can help the increased work potential of the T2S. An associ- ensure a higher operational efficiency of each stage. ated BSFC benefit of up to 12% in the low end and The presence of two units helps extend the high ef- around 1.5% in the peak region was observed. ficiency working range of the turbocharging system and avoids the need for titanium impellers as would Although not mentioned in the study [7], this addi- be required for high pressure ratio single-stage sys- tional work potential implies a lower enthalpy and tems due to excessive compressor exit temperatures hence temperature of the exhaust gas after the T2S [4]. which could have a negative impact on the exhaust aftertreatment (EAT) performance. This is owing to Other benefits include better transient and low-end re- the strong dependence of the conversion efficiency of sponse due to a smaller HP stage. Fuel consumption the EAT with exhaust gas temperature. While the im- benefits are attained from the lower backpressure due portance of this aspect has been highlighted in litera- to high efficiency operation of the two-stage system ture [3], it has not been considered within the primary when compared to a single-stage turbocharger for a scope of performance evaluation and appears to have given demand boost. With high boost pressures, two- received limited attention in previous studies. stage charging enables the use of more EGR while maintaining the air-fuel ratio (AFR). This helps reduce To summarise, two-stage turbocharging provides for NOX and improve the thermal efficiency [5]. better cylinder filling with a relatively lower backpressure penalty when compared to a single-stage tur- Two-stage systems provide operational flexibility with bocharging system. Although two-stage turbocharg- the use of by-pass valves and variable geometry ing provides this benefit, the more generally accepted turbine (VGTs) turbochargers. For example, the view with downsizing for heavy-duty application is the serial-sequential operation is where the HP stage is detrimental impact on component durability and re- matched for the low-end and transient needs while liability due to higher mechanical and thermal loads the LP stage is matched for the peak performance (higher PCP ∼30 bar greater than baseline) when requirements. The HP unit works in isolation until it considering the high mileage operation of vehicles in nears its choke limit after which the two units work this segment. The reduced cylinder approach is what in series with a by-pass valve directing high enthalpy is considered financially more attractive; although, it exhaust to the LP stage. Depending on the HP stage is not an obvious choice owing to customer accep- design, it can be completely phased out for high mass tance of reduced cylinder concepts and the treatment flows with only the LP stage providing the boost for the required for associated NVH issues [1]. rated condition. The mechanical limit quantified by the PCP does Typically, single-stage performance exceeds two- not allow for the practical adoption of two-stage tur- stage performance around the region of operating bocharging into heavy-duty engine application. This mode transition. This emphasises the need for a ro- implies that a complementary engine component or bust control system to reduce the detrimental mode system is required to utilise the potential benefits of transition effects in serial-sequential setups [6]. two-stage turbocharging and upon reconsideration of potential technologies for future heavy-duty engines The system architecture for two-stage systems are from [1], VVT of the intake valve which could throttle more complicated than with single-stage turbocharg- the flow and reduce the base engine volumetric effi- ers. In addition to by-pass valves and VGTs for op- ciency appears worthy of consideration. erational mode flexibility, the architecture of the EGR loop (HP, LP or dual) and the presence (or absence) A consequence of VVT on the intake valve is a fun- of interstage coolers (cooler between the LP and HP damental change in the thermodynamic cycle of the compressor) will determine the systems capabilities engine that is referred to as the Miller cycle. While and requirements to meet a certain engine air de- its combination with two-stage turbocharging appears mand. Interstage coolers are typically not used when to provide an alternative approach to realise the fu- demand boost is below 3.7 bar for light-duty applica- ture demands of heavy-duty engines, a brief overview tions as the disadvantages with cost, pumping loss of the Miller cycle is provided followed by an under- and system complexity exceeds the efficiency benefit standing of previous studies on combining the Miller for favourable compressor operation [4]. cycle and two-stage turbocharging.

An experimental study on a light-commercial vehicle engine [7] involved an exergy analysis of the ex- The Miller Cycle The Miller and Atkinson cycles haust gas after the turbocharger between a single- are over-expansion cycles wherein the effective com- stage and two-stage turbocharging (T2S) system pression ratio of the engine is less than its expansion while maintaining baseline performance. Tests along ratio. This is realised through VVT by changing the the full load curve showed a 30-45% increase in ex- intake valve close (IVC) from baseline to early (EIVC- haust exergy to work conversion in the T2S case. Miller) or late (LIVC-Atkinson) closing. In the case The analysis also shows that the loss due to addi- of EIVC, the charge undergoes an expansion to lower tional components in the air system is overcome by temperatures and pressures followed by the compres-

14 sion stroke and when LIVC is followed, a portion of the creasing its pressure through the charging system is charge is expelled out of the cylinder due to extended one of the primary uses of the Miller cycle [9]. intake valve open during compression. In both cases, the cylinder volume at IVC is the determining factor For diesel engines, Miller timing does not necessar- of the effect and the temperature and pressure at the ily lead to a decrease in NOX. This is largely due end of compression is less than baseline for the same to the nature of diesel combustion with a pre-mixed charge mass in the cylinder [8]. phase and a diffusion (mixing controlled) phase. Nor- mally Miller timing leads to a deterioration in com- Schutting et al [9] characterise the fundamental appli- bustion conditions and increases the ignition delay cations of the Miller cycle to be: period. Under extreme Miller cases and low loads, the ignition delay period is significant enough for the pre-mixed combustion portion to dominate and con- • throttling (alternative for conventional throttle) tribute to equal or higher NOX levels when compared to baseline IVC settings. • reducing the effective compression ratio This trend is less pronounced over high loads due to • outsourcing compression from the cylinder with better ignition conditions and longer duration of the charge cooling possibility diffusion phase with lower heat release due to low re- actant temperatures [12]. Under part-loads, so long All recorded uses of the Miller cycle are said to be de- as the mixing controlled portion is significant in com- rived from these fundamental attributes and are listed parison to the premixed portion, an increase in the under this section to provide greater qualitative and premixed peak would lead to a reduction in NOX ow- quantitative details on Miller evaluation methods and ing to its occurrence in fuel rich regions wherein NOX its impact on the combustion process and related en- formation would be limited [13]. gine metrics. Use of EGR is an added advantage in maintaining the Schutting et al [9] also explain that the over-expansion low temperature benefit with Miller timing. While Miller effect when attained through the use of VVT of the in- timing reduces the process temperature and NOX for- take valve is to be credited to Ralph Miller (Miller cy- mation tendency, it is not viable to act as a complete cle) while when attained through changes in the me- substitute for EGR when concerning NOX reduction chanical linkages of the cranktrain should be credited [8]. to James Atkinson (Atkinson cycle). The discrepancy was rooted in the fact that Miller had proposed his Emission formation appears to be affected by the na- cycle considering only EIVC and in 1982 David Luria ture of in-cylinder flow with Miller timing especially for realised the Atkinson cycle by using LIVC instead of the EIVC case. When EIVC was applied, the tur- a complex crank mechanism and termed it the Otto- bulent kinetic energy and local velocity in the piston Atkinson process. bowl at top dead centre (TDC) was found to reduce significantly. This factor is cited as a possible reason The realised reduction in the in-cylinder temperature for an observed increase in soot, hydrocarbons (HC) and carbon monoxide (CO) when EIVC was used for leads to a reduction in NOX since its formation is strongly coupled to mean cylinder temperature [10]. NOXreduction [8] [14]. On the contrary, EIVC was Lower temperatures also lead to reduced heat losses shown to reduce soot formation due to better mixing and improved working fluid properties that enhance and presence of fewer fuel rich regions in the squish the thermal efficiency [11]. region of the piston bowl due to reduced occurrence of charge mass fraction that satisfied the temperature The change in IVC leads to a decrease in charge and equivalence ratio conditions conducive for soot mass in the cylinder as Miller timing produces a throt- formation [15]. tling effect and reduces volumetric efficiency. This must be compensated by increased boost pressure Apart from showing an improved break specific to maintain or increase the charge mass. This can fuel consumption (BSFC)-NOX trade-off between the be viewed as an outsourcing of compression to the Miller case and baseline, the study by Theißl et al [10] charging system from the cylinder. The degree of shows Miller timing to be a better alternative for ex- Miller timing is hence limited by the capability and effi- haust thermal management at low loads when com- ciency of the boost system and intercooler(s). There- pared to throttling. Selective catalytic reduction (SCR) fore, two-stage turbocharging can be viewed as an systems typically require the exhaust temperature to ◦ enabler for Miller timing to more extreme degrees due be greater than 200 C to maintain reasonable con- to its ability to provide higher boost pressures [8]. version efficiencies and the throttling effect of Miller was shown through simulations to increase turbine ◦ While the concept of reducing the cylinder filling effi- exit temperature by 100 C from baseline. This was ciency with a change in IVC and compensating it with achieved while maintaining a very favourable BSFC an advanced charging system may appear ridiculous, trade-off when compared to using a throttle. the possibility to cool the charge while retaining or in-

15 Kovács et al [16] highlighted the performance in- As analysing Miller timing demands a fully controllable crease potential made possible by the PCP margin at- valvetrain, impact of Miller timing on combustion is tained by using Miller timing. When the charge mass typically studied on single-cylinder research engines and AFR were maintained equal to baseline, the PCP and can be considered as the traditional means of attained for the same performance was around 10 evaluating Miller timing [8] [14] [16]. The boost pres- bar lower than baseline. This PCP margin could be sure is provided by an external compressor and can used to increase the performance by increasing the be independently controlled while the backpressure fuelling, air pressure or phasing the heat release rate can be set based on a pressure regulating valve in (HRR) through advancing CA50 towards TDC. the exhaust. Evaluations are performed by setting the intake and exhaust pressures equal, or maintaining Malin et al [13] studied the Miller effect in combina- a constant pressure differential to aid HP EGR [14]. tion with swirl on the combustion and emissions of a In some studies, the baseline differential pressure is heavy-duty single-cylinder research engine with EIVC retained with the assumption of a higher efficiency upto 70◦CA (crank angle) and LIVC upto 100◦CA. charging system for the full engine [16]. To have more Tested cases were in the part-load region with and realistic estimates, boundary conditions from high fi- without EGR and an additional case at low load with delity full engine models would help better understand swirl variation by forcing the same mass flow through the interactive effects of Miller timing and the charging a single intake valve. This was performed while main- system. taining the mass flow and combustion phasing consistent to baseline IVC. Two-Stage Turbocharging with Miller Al- It was observed that the combined effect of Miller though two-stage turbocharging was in use since the timing and EGR was more significant on reducing 1970s on marine and commercial diesel engines [5], the motoring pressure and temperature. Miller tim- interest in its application for light and heavy-duty on- ing resulted in a small NOX reduction without fuel road application has renewed only in the recent past. consumption penalties and the exhaust temperature of non-EGR cases showed an increase. EIVC typ- Millo et al [18] performed an investigation into the po- ically showed a greater pumping loss increase than tential of two-stage turbocharging with the Miller cycle LIVC. Increased swirl lead to greater charge temper- by performing simulations on a detailed and validated atures and heat loss along with greater pumping loss heavy-duty engine model at one operating point (1500 and fuel consumption due to flow being channelled RPM 100% load). The T2S did not use an interstage through a single valve [13]. cooler. When applying the T2S without Miller timing, a maximum break mean effective pressure (BMEP) Zhang et al [17] evaluated the late Miller timing po- increase of 3% with peak pressures exceeding base- tential on reducing thermal and mechanical loads at a line (160 bar) was observed even when reducing the high speed and load condition of a heavy-duty single- compression ratio (from 17:1 to 16:1) and imposing cylinder engine. LIVC upto 50◦CA was instituted us- fuelling limits. A BSFC reduction of 2% was achieved ing three camshafts with different timing and tested although NOX increased by 10%. When coupled with at 25 bar indicated mean effective pressure (IMEP) a suitable Miller timing (EIVC), the BMEP increase and 3600 RPM with lambda as the swept parameter achievable was 5-6% for a marginally better BSFC re- and combustion phasing being maintained consistent duction of over 2% and a NOX reduction of 10% from to baseline IVC. the baseline due to reduced in-cylinder temperature.

The Miller effect lead to a significant reduction in ef- A limited number of experimental studies on Miller fective compression ratio from around 11 to 5 along timing were performed on serial engines. Ojeda [15] with the peak cylinder pressure reducing by a maxi- studied the effects of EIVC upto 100◦CA from base- mum of 23 bar. The maximum rate of pressure rise line on diesel combustion of a medium-duty truck en- could be reduced by increasing lambda greater than gine having a T2S and HP EGR. The EIVC was en- the baseline value of 1.6. Peak cylinder tempera- abled by a electro-hydraulic variable valve actuation ture and exhaust temperature dropped by 240 K and (VVA) device and the study showed lower fuel con- 70 K respectively at the baseline lambda setting. A sumption (upto 5%) and soot (at <5 bar BMEP load) lower lambda lead to higher peak and exhaust tem- for a constant NOX target. Exhaust temperature in- peratures. The Miller loss was quantified as poten- creased by 100◦C and soot reduction was attributed tial compression lost and showed a 11 fold increase to better mixing. An important observation from this from 0.1% of the indicated work to 1.13% for the max- study was the need for individual cylinder combustion imum LIVC. While a maximum benefit of 4.6% on fuel phasing control to maintain combustion stability be- consumption and nearly 50% NOX reduction was ob- tween cylinders. served at baseline lambda with LIVC, the smoke number showed an increase for the more extreme LIVC While the potential for improvements in engine perfor- cases. NOX reduction was presented in PPM and mance and emissions clearly exists with Miller and hence is subject to scrutiny given that lambda (air two-stage turbocharging, very few publications ad- mass flow) was swept through the study [17]. dress this issue for heavy-duty engines. The typical

16 evaluation method involves one of the approaches The benefits and impact of Miller timing are largely mentioned previously under “The Miller Cycle” sub- defined on a similar basis as the reference [16] and section wherein studies are performed on a single- hence it would be interesting to understand the poten- cylinder configuration of the engine and the boost sys- tial deviation in response parameters from the base- tem requirements are not represented during evalua- line between Miller evaluation on a single-cylinder tion. While single-cylinder studies are warranted for with an independent charging system (external com- studies focusing on understanding the isolated Miller pressor) and that of a full engine with a T2S. The in- impact on combustion, the intake and exhaust sys- take and exhaust systems are inherently coupled in tems are decoupled and not subject to the restric- the full engine case and the benefits are limited by the tions imposed by the turbocharging system. Addi- capability of the turbochargers. This point was specif- tionally, potential cylinder-cylinder interactions which ically mentioned in the reference but no publication till would represent the serial engine case would not be date appears to address this limitation explicitly due captured in the single-cylinder approach. to the focus placed on studying the Miller cycle in isolation. Another aspect that has received mention in literature but not included in Miller evaluations is assessment Additionally, the interaction between cylinders in case of the impact on EAT inlet temperature when using of a serial engine with LIVC are not accounted for in a T2S boosting system. The constraint this param- the single-cylinder approach. A realistic charging sys- eter imposes will largely depend on the temperature tem on a full engine with the Miller cycle such as a sensitivity of the SCR and the conversion efficiency T2S could potentially lower the EAT inlet temperature to be maintained with baseline in relation to the mag- as mentioned earlier. The significance of these as- nitude of reduction of exhaust gas temperature. This pects are also investigated in this study. disconnect between the benefits reported based on isolated Miller evaluation and the coupled Miller-T2S RESEARCH QUESTIONS evaluation considering the turbocharging system impact and the cylinder interactions in a serial engine The following research questions were formulated forms the basis for this study. and answered through the course of the study: PROBLEM DESCRIPTION 1. What is the maximum performance enhance- The focus of the study was to contrast Miller cycle ment possible with Miller timing evaluation on: evaluation between a single-cylinder and serial engine configuration. The serial engine utilises a two- (a) a single-cylinder basis limited by the case stage turbocharging system (T2S) as an enabler to constraints? meet the air demands when using the Miller cycle. (b) a serial engine with a T2S limited by the case constraints? The evaluation places emphasis on the performance increase potential with the Miller cycle. Parameters of 2. What is the deviation in performance enhance- interest include the Mean Effective Pressure (MEP), ment when comparing Miller evaluation based on the thermal efficiency and the specific fuel consump- the single-cylinder approach and serial engine tion while maintaining the specific NO emissions to X with T2S? baseline. For the serial engine, estimates of the boost system requirements when considering a T2S are 3. What are the turbocharging system demands to made in addition to the primary evaluation goals to meet engine performance objectives with Miller take account of the charge air requirements. timing? The study is performed at one engine operating point 4. What is the impact of the T2S on EAT inlet tem- which is of significance to the target performance for perature and SCR conversion efficiency? the T2S. The baseline is a six-cylinder HDD engine with a single-stage VGT turbocharger and HP EGR. The study is motivated by a recent publication on the Miller cycle based on single-cylinder experiments and METHODOLOGY simulations on a HDD engine which showed a 5-10% performance improvement (IMEP) from reference set- Traditionally, the Miller cycle evaluation is performed ting (without Miller) at a full load point depending on on single-cylinder engine test beds along with a cal- the IVC timing. The basis for this study will be to ibrated single-cylinder model on a one-dimensional adopt the approach and considerations used in the engine performance tool. As the study aims to esti- reference [16] to produce the traditional evaluation of mate and evaluate the Miller cycle on a single-cylinder the Miller cycle from a single-cylinder equivalent of the and full engine with a T2S based on an existing sim- baseline engine. ulation model as baseline, the study was proposed to be carried out purely based on simulations in GT- Power.

17 The study focused on comparing the single-cylinder Basis of Miller Evaluation The validated Miller evaluation and the coupled Miller-T2S evalua- baseline model does not include emission measure- tion on a full engine model at one engine operating ments or predictive combustion capability and hence point (1000 RPM and 100% load) which was selected was updated with predictive combustion capability as a point of signiﬁcance for the turbocharging sys- and run to estimate baseline NOX emissions. Re- tem. The signiﬁcance lies in the turbine work share sponse parameters from the baseline engine model between the two stages and this point is normally as- were crucial to determine the target values for the pa- sociated with the phasing-in of the LP stage as was rameters that were held consistent over the Miller cy- observed in a comparable HDD engine with Miller tim- cle evaluation cases. ing and T2S [19]. The Miller cycle was implemented by changing the The sequence of tasks performed is depicted in Fig- IVC from baseline value to late and early closing of ure 1 and is followed by a list of steps covered in the intake valve. LIVC conditions were implemented greater detail under this section. by imposing a dwell period at peak lift of the valve while EIVC conditions were based on scaling of the lift to maintain the same valve acceleration as baseline. Three EIVC and LIVC conditions were evaluated within a range of ± 90◦ CA from baseline IVC.

To ensure that the results are comparable between the baseline and Miller cases on the single-cylinder and serial engine with T2S, the following parameters were maintained equal to the baseline value as used in the reference [16]:

• Cylinder charge mass (air + EGR)

• Speciﬁc NOX • Crank angle at 50% burned fuel mass (CA50) Figure 1: Schematic of adopted methodology with start of injection (SOI) adjustment

Predictive Diesel Combustion Model . . . . . 18 • Scavenging pressure • Constant engine inlet temperature of 40◦C Basis of Miller Evaluation ...... 18 • PCP limited to baseline value Single-Cylinder Miller Evaluation ...... 18

Serial Engine-T2S Miller Evaluation . . . . . 19 Cylinder averaged values were used for the total cylinder charge mass and CA50 while the PCP target was Limitations ...... 19 set based on the observed value in the single-cylinder model without Miller (baseline IVC).

Predictive Diesel Combustion Model The study necessitates the model to have predictive Single-Cylinder Miller Evaluation The combustion capability as Miller timing and the corre- single-cylinder model consisted of one cylinder from sponding boost will change the start of combustion the baseline engine model with most of the air system conditions. The base model utilised fixed burn rates piping retained. The boost pressure was provided based on in-cylinder pressure measurements at by modelling an independent compressor and the 19 full load points of the base engine. Imposed exhaust pressure was regulated by an orifice with a burn rates are insensitive to changes in the start of variable diameter to maintain the scavenging pres- combustion conditions and hence not applicable to sure to baseline target. Cooled EGR was modelled cases such as Miller timing evaluation. as a LP circuit after the backpressure orifice and was regulated to match baseline specific NOX. The predictive combustion model calibration was performed on a single-cylinder model of the base engine The performance increase was analysed subject to: with the in-cylinder conditions imposed from the base engine conditions [20] for 10 full load points. Upon determining the calibration factors, the predictive ca- • increasing injected fuel mass pability was verified over 5 other full load points. The • increasing charge air mass calibration process is explained in greater detail under the engine modelling section that follows. • increasing geometric compression ratio

18 The evaluation included a base case with constant – Boost system estimates are subject to tar- ISNO and Miller timing as the sweep parameter and get efficiency and wastegate logic additionally a case with constant EGR rate to isolate – Compressor and turbine efficiencies set at the Miller effect on NOX reduction. 0.80 [19] – Interstage cooler was fixed in the air system architecture Serial Engine-T2S Miller Evaluation The serial engine model utilised six single-cylinder mod- – Charge air cooler outlet temperatures were els along with the baseline intake and exhaust man- fixed ifolds. The turbochargers were represented by the – Valve timing effects on in-cylinder flow could ideal compression and expansion processes with the not be resolved due to reduced order simu- best turbocharger stage efficiencies (∼64%) stated in lation literature [19]. – Ideal representation of turbochargers – Exhaust valve timing and profile was held The T2S was modelled as two compressors and ori- consistent with baseline (no changes) fices that correspond to the turbines. Bypass valves were substituted with wastegates for both the HP and • Engine Mechanicals LP stage to relieve the HP stage from its choking con- – PCP limited by baseline value and not de- dition and for the LP stage to meet the demand boost. sign limit The LP turbine was connected in series with no by- – Auxiliary loads like oil pump, power required pass to reduce the two-stage operating modes. to actuate intake valves under VVT and CAC pumps was not be considered The pumping work is dependent on the turbocharging system efficiency and performance demand based – Intake valve profile for LIVC and EIVC were on the IVC condition. This is expected to limit the derived from base model benefits observed in the single-cylinder Miller evalua- • Modelling tion which assumes a constant scavenging pressure equal to baseline implying higher charging system ef- – Modelling accuracy is dependent on accu- ficiency. Impact of the T2S on the EAT inlet tempera- racy of baseline validated model ture and subsequently on SCR conversion efficiency – Emission estimate was based on thermody- was also studied. namic in-cylinder conditions and was not be calibrated against experimental data due to Another aspect that was analysed was the cylinder- its non-availability cylinder interaction with Miller timing. This was ex- – Heat loss models were based on validated pected to be an important aspect to consider and the baseline model impact on response parameters on a cylinder-cylinder – Developed models require experimental val- variation basis was evaluated in this study. idation • Resources Limitations The study was bounded by the fol- – Only simulation based study was possible lowing limitations: due to non-availability of test facilities

• Engine operation ENGINE MODELLING APPROACH AND – Only steady state engine operation was CONSIDERATIONS considered – Evaluation and estimation was limited to The modelling considerations are presented in the fol- one engine operating point at full load lowing sequence: – Pilot injection was not used as a control pa- Introduction to GT-Power ...... 20 rameter – HRR phasing from baseline was not evalu- Base Engine Description ...... 20 ated – EAT inlet temperature impact was analysed Base Engine Model Capabilities ...... 20 purely in terms of temperature-conversion efﬁciency relationship Predictive Combustion Model Calibration . . 21 – Only diesel fuel was evaluated Predictive NOX Model Calibration ...... 26

• Air System Single-Cylinder Miller Evaluation ...... 27 – EGR was modelled only as an LP circuit Serial Engine Miller Evaluation with T2S . . . 29

19 Introduction to GT-Power GT-Power is a (<1000RPM) which have a pilot injection in addition module under the software package GT-Suite that is to the main injection. The presence of a pilot injection widely used by engineers and researchers for per- was ascertained by viewing the burn rate for these forming engine and vehicle simulation studies. GT- cases like in Figure 2. The significance of not having Power is used for engine performance studies and is this data for the study is explained under the predic- based on one-dimensional fluid dynamics to account tive combustion model calibration section that follows. for flow and heat transfer effects in the engine system. Model building is object oriented and the software includes a post-processing module GT-Post. GT-Power has the ability to perform predictive or non-predictive analysis and can be used for assessing steady-state, transient and controller performance of the engine [20].

Base Engine Description The base engine details are described under Table 1.

Table 1: Base Engine Specification Engine Type Inline-Six HDD Euro V Figure 2: Imposed burn rate from base engine model Bore 130 mm at 800RPM with two peaks indicating the pilot and Stroke 160 mm main injections Cylinder Displacement 2124 cm3 Number of Cylinders 6 The “flow heat transfer model” was used to account Total Displacement 12742 cm3 for the cylinder heat transfer effects by determining Compression Ratio 17.3 the convective heat transfer coefficient. This model Number of Valves 4 takes into account the in-cylinder swirl, tumble and Charging Single-Stage VGT turbulence which are computed based on geometric details of the piston bowl [20]. EGR HP Cooled EGR Maximum Power 353 kW @ 1600 RPM To account for friction effects, a constant friction load Maximum Torque 2541 Nm @ 1300 RPM of 5 Nm was modelled to represent the cooling fan Maximum BMEP 25 bar load in addition to the “Chen-Flynn model”. This Max. Cylinder Pressure 200 bar model is an empirical estimation of engine friction as Max. Boost Pressure 3.74 bar friction mean effective pressure (FMEP) and is a fac- Max. Boost Temperature 476 K tor of maximum cylinder pressure, mean piston speed Minimum BSFC 190 g/kW-h and square of mean piston speed [20]. Only the Maximum EGR Rate 28.8% square of mean piston speed factor was used in the base model and the mechanical efficiency observed at the evaluation point was 0.98. Base Engine Model Capabilities The base engine model was validated for the performance of 19 The valves were modelled with CA resolved lift pro- full load operating points covering the engine speed files along with the swirl coefficient and the forward range (800RPM-2300RPM). Emission measurements and reverse discharge coefficients with respect to were not incorporated into the base model. lift/diameter (L/D) for the intake valves. The exhaust valves had lookup tables for the forward and reverse The model was made non-predictive implying that the discharge coefficients as a function of L/D and static- burn rate is imposed from in-cylinder pressure mea- static pressure ratio across the valve. surements. GT-Power has the ability to determine the apparent burn rate from measured in-cylinder pres- The model included controllers for EGR rate, VGT sure traces with the objective of matching the CA re- rack position and injected fuel mass/cycle with target solved pressure estimate to measurement. Fuel in- values entered in tables with respect to engine speed. jectors were modelled with nozzle geometry details All piping including the coolers for the compressor and and total injected mass/cycle for each point. GT- EGR computed thermal effects to determine element Power estimates the demanded injected mass/◦CA in and gas temperature. Pressure loss along piping was order to match the imposed burn rate from SOI which accounted for purely by the viscous effects in the gov- was also defined as an input. erning fluid dynamic equations with no multipliers and no loss along bends and tapers. This is a modelling While the advantage of this approach is the reduced approach to avoid overestimating flow friction as dis- computational time, details of CA resolved injected charge coefficients at the valves typically include the mass and pressure profiles were not available. This total loss along the flow path (ports) and not in isola- was critical for cases such as the 4 low speed points tion.

20 Muffler backpressure was modelled as an orifice • Premixed Combustion Rate Multiplier (PCRM) downstream of the turbine. Turbocharger maps were used to determine the compressor and turbine oper- • Diffusion Combustion Rate Multiplier (DCRM) ating points while the turbocharger speed was set as an input for each operating point. The calibration of the model with measured datasets is required to provide predictions that are physically Predictive Combustion Model Calibration relevant of the reference case. While it is recom- The study necessitates the model to have predictive mended to calibrate the predictive model over a mini- combustion capability as Miller timing and the cor- mum of 25 points [20] to cover the full engine operat- responding boost will change the start of combus- ing range, the current study is limited to the available tion conditions. The base model utilised fixed burn data points (19 full load points) and intends to evalu- rates based on in-cylinder pressure measurements ate only one full load point. as mentioned under the previous section. These imposed burn rates are independent of the combustion An extremely important aspect while utilising the pre- conditions and hence will not represent the actual na- dictive combustion model is the need for injector rate ture of combustion for the particular case when Miller profiles either with injected mass or with injection timing is analysed. Figure 3 illustrates the imposed pressure resolved with CA. As the base model did burn rate in the base engine model at the evaluation not have this information, the rate profile was ob- point. tained from another study on the same engine [22] under a different output rating. The profile (Figure 4) was provided for one operating point (1200RPM and 1850Nm) in terms of injected mass with CA and was measured in an injector rate tube.

To ensure that the referred injection profile was physical for the conditions under investigation, a study on injector rate shaping on medium duty diesel engines [23] was referred to. This study provided a physical basis for instantaneous fuel mass flow rates through the injector and injection duration for the main injection. The obtained injection profile appears to follow a ramp profile with injected mass flow rate increasing with ◦CA as can be seen in Figure 4. The reference Figure 3: Imposed burn rate from base engine model profile had a total injected mass of 182 mg/cycle with at 1000RPM an injection duration of 16.975 ◦CA. GT-Power has two predictive combustion models for diesel combustion, DIPulse and DIJet. DIPulse is documented to provide similar or more accurate predictions than the DIJet model along with lower com- putation time [20] and hence the former was chosen as the predictive combustion model for the study.

DIPulse is a combustion model which tracks individual injections as pulses wherein the cylinder is discretized into 3 distinct thermodynamic zones: the main unburned zone which encompasses all trapped mass in the cylinder at IVC, the spray unburned zone with injected fuel and entrained gas and the spray burned zone with the combustion products [20] [21].

The model accounts for the different phases of combustion from fuel injection including spray entrainment, fuel evaporation, ignition delay and premixed and diffusion controlled combustion through physical sub-models. The model can be calibrated by 4 multi- Figure 4: Reference injection proﬁle [22] pliers for the mentioned phases of combustion:

The adoption of this injection proﬁle into cases un- • Entrainment Rate Multiplier (ERM) der study required scaling of the proﬁle with respect to injection duration. GT-Power automatically scales • Ignition Delay Multiplier (IDM) the instantaneous injection mass with respect to the

21 total injected mass and hence this aspect was inconsequential. A basis for injector duration scaling was to maintain physically acceptable injection pressures. As the evaluation was performed at only one engine operating point with a ﬁxed injection rate proﬁle, the scaling was performed with respect to the evaluated reference case (1000RPM full load).

The case required the injection duration to be increased by 5.5 ◦CA to 22.475 ◦CA to maintain acceptable injection pressures. The average injection pressure under the scaled profile was ∼1600 bar and the maximum injection pressure was ∼2050 bar. The Figure 7: Comparison of injection pressure profiles maximum acceptable pressure was referred to from the base model. Figure 5 illustrates the injector mass flow and pressure profiles for the 1000RPM case. include:

• The profile may not be representative of the actual injection profile for the evaluated case (1000RPM full load) under a different output rating • Combustion calibration multipliers obtained with this injection profile are valid only with this injection profile

Under these circumstances, the calibration was performed over 10 operating points with 5 points purely for validation (Figure 8) of the predictive model capa- Figure 5: Scaled injection pressure and mass profile bility. The 4 low speed points (<1000RPM) mentioned for 1000RPM case earlier were not used due to lack of data regarding pilot injection profile and the injected mass. This was to The injection profiles for other operating points in the avoid modelling them based on assumptions and in- calibration data set were scaled for injection dura- fluencing the calibration parameters for the evaluation tion with respect to the evaluation case (1000RPM). case which does not include pilot injections. The scaling was performed to provide similar instantaneous fuel flow rates and injection pressures while varying the duration to match the total injected mass. This was performed with multipliers that normalised the engine speed and total injected mass for the operating point with respect to the evaluation point. Fig- ure 6 and Figure 7 compare the scaling of injection profile with mass flow and pressure for the cases at 1000RPM, 1600RPM and 2300RPM.

Figure 8: Engine operating points for calibration and validation of the DIPulse

The calibration multipliers were determined on a single-cylinder model of the base engine under what is termed as cylinder pressure only analysis (CPOA) with a genetic algorithm (GA) optimiser scheme. CPOA evaluates the predictive combustion model in relation to the measured in-cylinder pressure trace Figure 6: Comparison of injection mass proﬁles between IVC and exhaust valve open (EVO) and hence requires boundary conditions that are descrip- The injection duration for the other cases varied be- tive of the trapped in-cylinder conditions at IVC (cycle tween 24.63 ◦CA and 32.23 ◦CA. Adopting an injec- start). The model consists of the injector, cylinder and tion rate proﬁle for the study has its limitations which crankshaft objects as can be seen in Figure 9.

22 prediction’s insensitivity to the changes in IDM and PCRM could be biased by having performed the calibration only at operating points under full load where the ignition conditions are more favourable. These two parameters were observed in reference studies to have more impact on pilot injections even under full load conditions [21].

To ensure that the obtained multipliers would be physically relevant, the combustion model’s predictive capability was evaluated in accordance to a study that focused on this aspect. The study [21] provides a reference for acceptable predictive capability of a calibrated DIPulse model with respect to BSFC and BMEP deviations over the full load operating range.

The BMEP and BSFC predictions in comparison to the baseline non-predictive model are illustrated in Figures 10, 11, 12 and 13.

Figure 9: CPOA model in GT-Power

The parameters used to initialise the cases can be referred in Figure 71 under the appendix. The initialisation parameter values were obtained as results from the baseline model.

The GA optimiser settings were ﬁxed based on guidelines from GT which were documented to provide satisfactory results for calibration of the DIPulse combustion model [20]. The objective function in the opti- Figure 10: Predicted BMEP at Full Load miser was set to minimise the “Improved RMS Burn Rate Error” which represents the deviation between the burn rate from the predictive combustion model and the non-predictive one based on measured pressure. The range of DIPulse calibration multipliers is provided in Table 2.

Table 2: Recommended Combustion Calibration Mul- tiplier Settings for GA Optimiser ERM 0.95 - 2.80 IDM 0.30 - 1.70 PCRM 0.05 - 2.50 DCRM 0.40 - 1.40 Figure 11: Deviation of BMEP Prediction at Full Load The DIPulse model calibration was attempted with moderate changes to the optimiser settings wherein The maximum percentage deviation in BMEP and the IDM and PCRM ranges were extended but the op- BSFC was observed to be ∼3% and the average per- timisation results appeared to converge at the values centage deviation was ∼1.5%. This can be attributed listed in Table 3. to the deviation in pressure prediction which impacts the performance. These results compare well with [21] in which the BSFC predictions under full load Table 3: Optimised Combustion Calibration Multipliers showed a maximum deviation of 8.9% and an average ERM 1.23 deviation of 2.3%. It must be noted that the calibration IDM 0.30 in the study [21] was performed over the full engine PCRM 2.99 operating range with a more exhaustive data set to DCRM 1.39 represent varying combustion conditions. The predictive capability appears satisfactory within the available The IDM and PCRM values were the ﬁrst to reach the dataset as the validation points are within the maxi- recommended limits under the optimisation run. The mum deviations observed.

23 2300RPM. Both the agreement and deviation of predicted burn rate profile could be attributed to the fixed injection profile used in this evaluation and the use of one set of calibration parameters to represent different operating conditions.

To provide a basis for the observed deviation in burn rate phasing, a sensitivity analysis with respect to the IDM was performed over the 2300RPM case which was observed to have the maximum deviation in burn rate between predicted and baseline. It was suspected that the IDM was low for the high speed points Figure 12: Predicted BSFC at Full Load and hence the burn rates were phased earlier than baseline. The IDM value as a result of the calibration was at the lower end of the recommended range (0.30). As the multiplier directly scales the ignition delay, the sensitivity analysis was performed for IDM values of 0.40, 0.65 and 1 while retaining other multipliers to the calibrated value.

The sensitivity analysis showed that an increase in IDM increased the ignition delay by ∼4 ◦CA and had a stronger premixed portion of combustion that lead to the phasing of the burn rate earlier than the predicted burn rate with calibrated IDM as can be seen in Figure 14. Figure 13: Deviation of BSFC Prediction at Full Load The sensitivity analysis was inconclusive in determin- Additionally, the prediction of the in-cylinder pressure ing the cause for phase shift of the burn rate between and burn rates in comparison to the non-predictive prediction and baseline at higher speed points. Plau- base model were also observed to understand the sible factors could be the objective function in the cal- magnitude of deviation in this aspect which will impact ibration (only burn rate error RMS), number of points emission prediction in addition to performance. Fig- used for calibration, the injection proﬁle used and in- ures 15 and 16 illustrate the comparison at 1000RPM, herent errors in the model. 1300RPM, 1950RPM and 2300RPM. The points were chosen to represent the evaluation point and the points with maximum BMEP/BSFC prediction deviations (1300RPM and 2300RPM) along with the point showing the least deviation (1950RPM).

In-cylinder pressure traces are found to follow the measured pressure trends and typically overestimates the peak pressure at higher engine speeds and underestimates peak pressure at lower engine speeds. The maximum deviation in peak pressure estimation is over 7.5% with an average deviation of ∼4%.

When observing the predicted burn rate with the ap- Figure 14: Burn Rate Sensitivity with IDM at parent burn rate from the base model, two aspects 2300RPM deserve consideration:

1. Proﬁle of the predicted burn rate

2. Phase of the predicted burn rate

In Figure 16 it can be observed that the burn rate proﬁle and phase with the apparent burn rate from base model are in good agreement with the cases at 1000RPM and 1300RPM. A phase shift can be observed in the proﬁles of the cases at 1950RPM and

24 Figure 15: In-Cylinder Pressure Prediction vs Baseline

Figure 16: Burn Rate Prediction vs Baseline

25 Predictive NOX Model Calibration Once limitation of 1500 bar [25] in relation to engine param- the performance prediction capability of the combus- eters in the baseline predictive model. The operating tion model was ascertained to be acceptable, the conditions differed from the baseline and hence a di- next step within combustion model calibration was for rect correlation of calibrated combustion conditions to emissions, in particular NOX. NO is considered the NO estimates was not possible. major constituent of engine out NOX for diesel combustion and hence the options in GT-Power primarily The objective was redefined to understand the de- cater to NO prediction and possibly NO2 if specified viation in NO estimates between standard reaction explicitly by the user. rates (multipliers=1) and calibrated ones under the extended Zeldovich mechanism to use the former while By default, NO formation is estimated by equilibrium acknowledging its limitations. It was also decided chemistry which performs species concentration com- to avoid inclusion of other exhaust emission compo- putation of 11 products of combustion species in the nents (HC, CO and PM) due to lack of measured data lumped burned zone at the zone temperature and to calibrate these emission models. pressure [20]. This is computed for each ◦CA and the final value is obtained at EVO. The possibility of utilising the measured data for NOX calibration was explored by first comparing the in- The other option is to incorporate the chemical ki- cylinder pressure traces of the predictive combustion netics based extended Zeldovich mechanism for NO model and the cases presented in the reference [25]. prediction and calibrate the estimate with test data Figure 17 shows the in-cylinder pressure measure- through calibration multipliers to scale the reaction ments from the reference for two operating points at rates and the total NO estimated. NO formation is 1200RPM with an injected fuel mass of 135mg/cycle highly sensitive to the burned zone temperature and (high load) and 23mg/cycle (low load) without EGR. does not reach chemical equilibrium through the combustion event. Additionally, the decrease in burned zone temperature with expansion beyond the peak pressure point and mixing with cooler air or burned gas freezes the NO chemistry [24] implying that a lower decomposition of NO occurs towards EVO.

This is captured by the extended Zeldovich mechanism and calibrated NOX predictions from literature were found to represent measurements with an average absolute deviation of 25-30% over a speed and load sweep [21] and overestimates NO by ∼5% at one point when using the standard reaction rate constants [24].

It was expected that the equilibrium chemistry approach for NO would deviate significantly from the true value due to its simplifying assumption of infi- nite reaction rates. Hence, it was decided to make use of the extended Zeldovich mechanism as the NOX model for NO prediction while being certain of the lim- Figure 17: Measured in-cylinder pressure traces [25] itation of the equilibrium chemistry approach. The SOI for this operating point was estimated from neighbouring operating points of the base model to It is understood that for varying engine operating con- be ∼3◦CA before TDC (at full load). The electrical ditions, the extended Zeldovich mechanism would re- signals for the injection profiles were provided but not quire measured data for calibrating the estimates. A usable as the physical injection profile although, it was major limitation was the lack of measured emission understood that the two points had varying injection data incorporated into the baseline model. Account- profiles compared to the scaled and calibrated profile ing for NOX is essential to proceed to Miller evaluation used in the predictive model. The predicted pressure and hence literature was referred to possibly obtain trace (Figure 18) appears to match the measured val- emission data from previous studies on a similar en- ues reasonably well as shown in Figure 17. gine. On the basis of comparable pressure profile prediction, it was considered to use the data points at One such study involved NOX measurements on 1200RPM and varying loads (Table 4) between the the single-cylinder configuration of the full engine at two cases previously mentioned to calibrate the NOX the internal combustion engine laboratory at KTH. model. The variation was in the intake pressure and The single-cylinder engine had a lower displacement the injected fuel mass. The predictive combustion (∼8%), a marginally lower compression ratio, lower model was run with a NOX object which predicted number of injector holes and a peak injection pressure NO subject to six multipliers within the extended Zel-

26 deviation of ∼30% with the ﬁrst point under 1600RPM alone showing nearly 100% deviation. The default multipliers show a prediction that follows the same trend as the calibrated NOX model with a nearly uniform gain of 1.65 for cases at 1600RPM and 1.58 for cases at 1900RPM over the calibrated estimates.

Figure 18: Predicted in-cylinder pressure traces to represent measurements from reference [25] dovich mechanism to provide maximum degrees of freedom for the calibration.

Table 4: Test Points for NOX Calibration [25] Figure 19: NO model validation with measured data Relative Inlet Pressure Injected Fuel and comparison with default multipliers (bar) (mg/cycle) 0 23 Therefore the limitations of using the NOX model with 0.4 23 default multipliers would be an overestimate of NO by 0.75 23 a nearly constant factor at a given speed point. The 0 135 sensitivity of the model to EGR was not evaluated due 0.4 135 to lack of relevant measured data. Since Miller evalu- 0.75 135 ation was performed only at 1000RPM, the predicted 0.3 84 NO values are expected to be biased by around same 1.4 57 margin but subject to EGR and in-cylinder tempera- 1.74 93 ture variations. Miller evaluation was carried out with the default multipliers under these limitations.

Upon calibration of the NOX model, a comparison between NO predictions from the calibrated and default multipliers was performed at 6 validation points [25] listed under Table 5 (1600RPM) and Table 6 (1900RPM).

Table 5: Test Points for NOX model validation at 1600RPM [25] Relative Inlet Pressure Injected Fuel (bar) (mg/cycle) 0.6 81 1 66 1.76 76 Figure 20: NO prediction: Zeldovich Vs equilibrium chemistry

The shortcoming of the equilibrium chemistry ap- Table 6: Test Points for NOX model validation at proach was evident when comparing NO predictions 1900RPM [25] at 1200RPM high load and low load [25] with the Relative Inlet Pressure Injected Fuel equilibrium chemistry approach and the extended Zel- (bar) (mg/cycle) dovich mechanism. Equilibrium chemistry fails to cap- 1.2 49 ture the freeze in NO formation through the expansion 1.5 68 stroke and in cases where the premixed combustion 1.9 83 was stronger such as the low load point (23mg/cycle injected fuel), equilibrium chemistry overestimates NO as shown in Figure 20. The comparison between the calibrated NOX model, measured data and NOX model with default multipliers is illustrated in Figure 19. The calibration multipli- Single-Cylinder Miller Evaluation To es- ers enable the model to predict NO with an average tablish a basis for evaluation of the Miller cycle,

27 a single-cylinder equivalent of the baseline engine vation for this lies in the baseline model having a very model was developed. As previously mentioned, high mechanical efficiency of 0.98 at the evaluation Miller studies are traditionally performed on single- point and the need to approximate the mechanical ef- cylinder setups to have control over engine bound- ficiency of a single-cylinder engine to a certain value ary conditions and more importantly the valve lift and (0.9 in [16]) to be representative of a full engines’ timing events. The single-cylinder model was derived friction. To avoid the discrepancies from the friction from the baseline engine model along with the piping effects, indicated specific metrics were used through for the manifolds and EGR. the study.

Changes in the air path included routing of EGR as Consistency between the baseline model and the a LP circuit, implementing an isentropic efﬁciency single-cylinder model at the evaluation point without based independent compressor and a backpressure Miller timing required implementation of the condi- valve to impose ﬂow restriction in the exhaust. The in- tions followed in [16] into the model. The single- take and exhaust are thus decoupled in such a setup cylinder model required controller development for the as is the case in single-cylinder test beds. The in- following attributes: tercooler was modelled as a controllable heat sink to regulate engine inlet temperature. • SOI to control CA50 Piping geometry required tuning to match the PMEP • PCP limits for Miller cases through and net IMEP to baseline at the evaluation point. Tun- ing was performed by adding dead volumes to the in- – injected fuel mass take and exhaust manifolds along with a buffer volume ahead of the backpressure valve for damping – air mass through compressor power exhaust pulses from a single-cylinder. A schematic – geometric compression ratio of the single-cylinder model is depicted in Figure 21. • EGR rate to control ISNOX – EGR valve controller (nested controller) to realise the demand EGR rate

• Compressor power to control total trapped mass in the cylinder

• Backpressure valve to control exhaust manifold pressure and pressure differential across the cylinder

• Intercooler heat rejection to maintain constant engine inlet temperature

Control over these attributes was necessary to isolate the Miller effect through the evaluation. The baseline response derived target values for the controllers are summarised in Table 7.

Table 7: Target for control parameters through Miller evaluation Figure 21: Single-cylinder model schematic on GT- Control Parameter Target Value Power Total trapped mass 7120 mg Parameters including the trapped air at cycle start (air Scavenging pressure -0.1 bar and EGR mass flow rates), specific NO and temper- X PCP 187.2 bar atures and pressures across the cylinder were compared with the baseline model at the evaluation point. ISNOX 3.32 g/kW.h CA50 14.41 ◦CA ATDC Typical deviations were observed to be of the order ◦ of 1% and were achieved by developing controllers Engine Inlet Temperature 40 C for the parameters of interest. Matching the baseline PMEP required the most tuning in the exhaust piping The controllers were developed under recommended parameters and a deviation within 1.5% was achieved guidelines from GT-Power to provide simple imple- and considered satisfactory. mentation of parameter control. The guideline was to characterise the system as a single order linear time It was decided to evaluate the output specific metrics invariant system. This implied that controllers were (specific NO for example) with respect to the net in- built as single order PIDs with a P and I gain. GT- dicated power in place of the brake power. The moti- Power provided a time domain based pole placement

28 method to represent the system response to a step input and choose desired settling time (aggressiveness) of the controlled response.

It must be noted that the controllers were developed for the speciﬁc purpose of steady state simulations and considered systems as having a single input single output (SISO). While this reduced the complexity of controller development, the interaction between controllers required phasing-in of controllers and ﬁl- tering of some feedback signals to avoid instability. This lead to slower closed loop response from the controller-plant system but provided a single-cylinder model that was fully controllable with respect to the parameters of interest. The runtime to convergence for one case was typically observed to be within 11 minutes.

The valve profiles under Miller timing were derived from the baseline intake valve profile. LIVC was achieved by extending the dwell duration at peak lift and EIVC was achieved by scaling the acceleration profile of the baseline intake valve to be maintained for reduced open duration as shown in Figure 22. The Figure 23: Full engine T2S model schematic on GT- Miller study was performed at an IVC of ± 30◦, ± 60◦ Power and ± 90◦ CA wherein the ‘+0 indicates LIVC and a 0 ‘− indicates EIVC. The HP compressor did not have a bypass while the turbine stages were modelled with wastegates as an A contingency point at LIVC 75 was included after ini- alternative to modelling bypass ducts and valves. GT- tial trials of the serial engine model due to suspected Power models the turbine and wastegate as an ori- non-convergence of the LIVC 90 case. The mod- fice wherein the mass flow through the turbine orifice elling issues were subsequently sorted yet the results will provide a theoretical power estimate based on the include LIVC 75 as an additional point through the isentropic efficiency and the wastegate orifice is a free single-cylinder and serial engine Miller evaluation. flow opening in the flow path.

Serial Engine Miller Evaluation with T2S To arrive at a suitable size for the turbines (orifice For the full engine with T2S, modifications had to be diameter), controllers were developed to enable the made on the air path of the base model. Two tur- turbine (orifice) diameters to be a variable. This cor- bines and compressors were modelled to represent responds to a T2S system with 2 VGT’s with waste- the HP and LP stage with imposed isentropic efficien- gates. This intermediary setup was used to define the cies. The compressors had a fixed total-total isen- (fixed) geometry of the turbines based on the bound- tropic efficiency of 0.8 and the turbines had a fixed ary conditions to be representative of the base engine isentropic efficiency of 0.8 (total-static for LP stage). model at the evaluation point. These boundary con- The efficiencies were derived from literature to be rep- ditions are the same ones used to match the single- resentative of a high efficiency T2S system with tur- cylinder model to baseline. bocharger efficiencies of over 60% [19]. The steady state values of the turbine orifice diame- Additional piping between the stages were built with ters to match baseline condition were 30 mm for the the inclusion of an interstage cooler between the com- HP stage and 41.25 mm for the LP stage. Through pressors. To avoid the complexities of T2S sizing and the model tuning trials for the full engine model, it was matching, it was decided to model the T2S system understood that the LP stage turbine power was insuf- wherein the HP stage provided a constant total pres- ficient to meet the trapped mass criterion for the LIVC sure ratio (TPR) of 2.4 (maximum output condition) at 90 case by a deviation of 10%. To match the baseline the baseline point (1000RPM, full load and no Miller conditions and reduce the deviation, the LP stage ori- timing) and the LP stage would provide the additional fice diameter was reduced to 40 mm which lead to the boost to meet the trapped mass condition. The TPR deviation to reduce to 1.5% below target. The model was referenced from literature and so was the T2S was not further tuned as controller convergence tar- operating logic [19]. A schematic of the full engine gets were set at 1% and the turbines were made fixed model with T2S is shown in Figure 23. geometry with 30 mm and 40 mm orifice diameters for the HP and LP turbine respectively.

The full engine with T2S was a more complex model

29 Figure 22: LIVC profiles along with the valve acceleration, velocity and lift for EIVC cases requiring a greater number of controllers to maintain parameters at desirable values. New controllers were Table 8: Target for control parameters through Miller developed for the following attributes: evaluation on full engine Control Parameter Target Value • HP wastegate diameter for HP stage TPR Total trapped mass 7120 mg PCP 187.2 bar • LP wastegate diameter for total trapped mass ISNO 3.32 g/kW.h (cylinder averaged feedback) X CA50 14.41 ◦CA ATDC • SOI to maintain CA50 for each cylinder (reduce Engine Inlet Temperature 40◦C combustion variability) HP Compressor T-Inlet 60◦C LP Compressor T-Inlet 40◦C • Cooler heat rejection to maintain inlet tempera- HP Compressor TPR 2.4 tures – EGR cooler for LP stage inlet – Interstage cooler for HP stage inlet was deemed necessary as combustion variability was observed to be significant when a cylinder averaged – Intercooler for engine inlet feedback was used. The approach of SOI control for • EGR rate to control ISNOX individual cylinders to reduce combustion variability was mentioned in literature [15]. – EGR valve controller (nested controller) to realise the demand EGR rate The coolers in the model had a geometrical basis • PCP control through injected fuel mass (cylinder from the base model to replicate reasonable flow averaged feedback) restriction. To quantify the heat rejection required to achieve controller targets on inlet temperatures, The controller targets for the full engine case are the coolers were modelled adiabatic with a controller summarised under Table 8. based negative heat source term. The negative heat source at steady state would quantify the cooler ca- The T2S logic was based on control of the waste- pacity (heat rejection) required to realise the control gate diameter of the two turbines. The HP waste- demand. gate would ensure a TPR of 2.4 from the HP stage and wastegate excess flow while the LP wastegate A combination of T2S and LP EGR circuit reduces would provide the required power to the LP compres- the PMEP by 44% from the baseline full engine. The sor to match the trapped mass condition of the cylin- scavenging pressure increases from -0.1 bar to +0.06 ders. Feedback for cylinder trapped mass was cylin- bar while the intake and exhaust pressures drop by der averaged. Individual CA50 control through SOI around 4% and 9% respectively. The net IMEP de-

30 viated by 1.5% and total mass flow (air+EGR) match crease potential by different means. Additionally, the the baseline well within a 1% deviation margin. Miller effect for a constant EGR rate was also studied for a better understanding of the isolated effect of the The positive scavenging pressure difference is an im- Miller cycle on ISNO. portant observation to consider when analysing the results. This scenario arises due to the EGR loop The results are presented in two parts for each eval- being modelled purely as a LP circuit. The bound- uation case. The first part within each case will high- ary conditions imposed on the single-cylinder model light the observations for the given treatment. This were derived from the base model with a HP EGR cir- includes impact on cuit wherein the backpressure was a combined result of the turbine efficiency and the pressure differential demanded by the EGR rate. The full engine model • control parameters would only provide the backpressure as a result of – total trapped mass, ISNO, backpressure, in- the T2S setup and boost demand thus biasing results let temperature and CA50 in favour of this configuration. – injected fuel mass, air mass flow, boost RESULTS pressure and geometric compression ratio for the performance increase cases The results are presented first for the evaluation of • thermodynamic and combustion parameters the Miller cycle on the single-cylinder model followed by the serial engine with the T2S. A summary of the – cylinder pressure and temperature, ignition result sections that follow is presented below: delay, maximum pressure rise rate, lambda, %residual gas The Miller Cycle on a Single-Cylinder En- • performance parameters gine Configuration ...... 31 – net IMEP, PMEP, ISFC, ITE, cylinder heat Miller Effect for Constant ISNO . . . . . 31 loss and swirl

Miller Effect for Constant EGR Rate . . 39 The second part would illustrate crank angle resolved Performance Enhancement by Increas- observations that typically help explain and better un- ing Air Mass ...... 42 derstand the trends observed in the ﬁrst part of the results. Performance Enhancement by Increas- The Miller effect under constant ISNO is considered ing Fuel Mass ...... 46 as the base case through the single-cylinder analysis and references regarding the similarity or deviation in Performance Enhancement by Increas- observations were made with this case. ing Compression Ratio ...... 50

The Miller Cycle on a Serial Engine Conﬁg- Miller Effect for Constant ISNO The observations uration with T2S ...... 50 for the Miller effect with constant ISNO are clustered as sets under Figures 24, 25, 26 and 27. Miller Effect for constant ISNO . . . . . 50 Figure 24 shows the effect of the Miller cycle on Miller and Cylinder-Cylinder Interaction 56 the dynamic (IVC) compression ratio (DCR) which shows a more pronounced reduction with LIVC than Cylinder-Cylinder Variation: Serial En- with EIVC. At an EIVC of 30 CAD, the DCR is ob- gine Base Case ...... 56 served to increase and this could possibly be due to the favourable cylinder ﬁlling tendency with mild EIVC Performance Enhancement by Increas- at a low speed point such as the evaluation point ing Fuel Mass ...... 57 (1000 RPM). The trend with the motoring (compression) pressure and temperature is similar to the DCR Cylinder-Cylinder Variation: Serial En- variation with IVC. A reduction of 8.7 bar and 56 K at gine Performance Case . . . . . 60 extreme EIVC (-90) along with a reduction of 16.6 bar and 106 K at extreme LIVC (90) are observable.

The Miller Cycle on a Single-Cylinder En- The trapped mass and inlet temperature appear to be gine Conﬁguration As mentioned under the under control as required. The increase in boost pres- methodology section, the single-cylinder evaluation of sure required to maintain cylinder ﬁlling constant is the Miller cycle involved considering the Miller effect clear (bottom left corner of 24). An additional 1.57 for constant ISNO followed by the performance in- bar (4.83 bar) and 2.74 bar (6 bar) increase in boost

31 pressure with a corresponding backpressure increase The CA resolved mass averaged temperatures are to maintain the negative scavenging pressure from presented in Figures 30 and 31 which better ex- baseline are observable. An interesting observation plain the observations cited previously. Around 0◦CA, was the increase in exhaust temperature by 70◦C the reduced motoring temperature is evident followed (596◦C) for EIVC and 83◦C (609◦C) for LIVC which is by a marginal reduction in the peak temperature. thought to be due to the higher operational pressures Around 180◦CA, the higher exhaust temperatures for of the cycle. the Miller cases is evident due to the reduced expansion (to lower cycle pressures). Also visible for the Figure 25 shows the impact on combustion related EIVC cases (31) is the reduced in-cylinder tempera- parameters. SOI adjustment within 0.2 CAD was nec- ture in the intake stroke with EIVC (towards 540◦CA). essary to maintain CA50 constant. Since injected mass is constant at this point, the end of injection The CA resolved burned zone temperature variation (EOI) is offset by the SOI change. A marginal in- in Figure 32 helps understand the Miller effect on NO crease (<0.2 CAD) in ignition delay is observable and formation. It can be observed that the sharp tem- explains a marginal decrease in burn duration (<1 perature peak resulting from premixed combustion CAD from baseline). PCP and peak temperature are shows a progression to lower temperatures with Miller reduced by 4 bar and 20 K (EIVC) and 8 bar and 40 for both EIVC and LIVC. The premixed peak burned K (LIVC). The maximum rate of pressure rise does zone temperature reduces by 109 K (LIVC) and 177 not change signiﬁcantly but interestingly reduces with K (EIVC) followed by an increase to a largely similar Miller when compared to baseline. burned zone temperature beyond 10◦CA.

The air-EGR relation is shown in Figure 26 in order The nearly constant slope in burned zone tempera- to maintain ISNO constant. As increasing Miller (IVC ture between 20◦CA and 50◦CA maintains a maxi- deviation) reduced the process temperature, a lower mum difference of ∼40 K (LIVC) and ∼20 K (EIVC). EGR dependence with higher degree of Miller is ex- As NO formation is highly sensitive to the tempera- pected. This is supported by the observation of an ture, this trend shows the Miller cycles’ positive con- increasing lambda with Miller and a corresponding tribution to reduced NO. decrease in %residual gas (includes EGR). Although the maximum burned zone temperature appears to The apparent heat release rates (AHRR) in Figure 33 be nearly constant, the CA resolved plots that follow illustrate the effects of the marginal increase in igni- will better illustrate the Miller benefit on lower process tion delay with Miller which causes the AHRR to ad- temperature. vance leading to the reduced burn duration. The good ignition conditions at a high load point could explain The Miller effect on performance is illustrated in Fig- the lack of significance in these deviations. ure 27. PMEP shows a constant deterioration with Miller with a greater pumping loss at extreme EIVC Figures 34 and 35 showcase the pumping loop and than LIVC. This can partly be attributed to the reduced bottom portion of the high pressure cycle for the EIVC lift profiles used for the EIVC cases highlighted earlier and LIVC cases. It can be observed that with increas- in Figure 22. Both IMEP and ISFC show marginal im- ing Miller the operating cycle pressure increases. The provements up to 60 CAD Miller timing The benefit Miller loss mentioned earlier can be visualised with is explained by the marginal ITE increase due to re- ease through these figures. duced heat loss and improved working property of the gases. In the case of EIVC, as the intake valve closes in advance, an expansion of the trapped charge takes Heat loss reduction of 2kW (EIVC) and 1kW (LIVC) place (distinctly visible for EIVC 90). This expansion are observable. The reduced heat loss potential with causes an increase in the pumping loop area which LIVC could be hindered by the higher swirl which in- translates to an increase in gas exchange loss. This creases the heat taken up by the charge. increase in gas exchange loss from the baseline is quantified as the Miller loss for EIVC [16]. The reduced PCP provides the performance improvement scope for same mechanical loading as baseline For LIVC cases, the intake valve closes father into the and the reduced peak temperature leads to lower heat compression stroke. An ideal compression would be if loss and improved efficiency. The drop in the ITE to- the IVC were to happen at bottom dead centre (BDC). wards the extreme Miller cases can be explained by This reduced compression potential from BDC would an increase in what is termed as the Miller loss. This lead to an increase in the compression loss. The will be explained in greater detail subsequently. baseline IVC setting already accounts for some compression loss due to its occurrence past BDC. This The in-cylinder pressure traces are illustrated in Fig- increase in compression loss with respect to baseline ures 28 and 29. The Miller effect on reduced motoring IVC is quantified as the Miller loss for LIVC [16]. and PCP are evident here. What is also observable but perhaps not evident is an increase in the cycle exhaust pressure.

32 Figure 24: Observations for constant ISNO:Set I

Figure 25: Observations for constant ISNO:Set II

33 Figure 26: Observations for constant ISNO:Set III

Figure 27: Observations for constant ISNO:Set IV

34 Figure 28: In-cylinder pressure trace with EIVC

Figure 29: In-cylinder pressure trace with LIVC

35 Figure 30: In-cylinder temperature trace with EIVC

Figure 31: In-cylinder temperature trace with LIVC

36 Figure 32: Burned zone temperature with EIVC and LIVC

Figure 33: Apparent heat release rate with EIVC and LIVC

37 Figure 34: Pumping loop with EIVC

Figure 35: Pumping loop with LIVC

38 Miller Effect for Constant EGR Rate Although not The impact on performance is visible in Figure 39. the focus of this study, the Miller cycles’ all round ben- While the trends are comparable to the base case, efits motivates a need to understand the emission the additional EGR causes a maximum reduction in (NO) reduction potential and the corresponding im- ITE to the order of 0.5% relative to base case ITE. pact on performance. This was performed by holding This impacts the IMEP and ISFC to the same mag- the EGR rate constant to the baseline value (17.58%) nitude. Heat loss is marginally reduced due to lower and allowing the ISNO to vary (reduce) as a result of in-cylinder temperatures but is inconsequential at im- the Miller effect. All other control parameters are con- proving the ITE. sistent with the base case. The ISNO reduction observed earlier is better under- The motoring pressure and temperature were found stood through the crank resolved burned zone tem- to be similar to the base case while the DCR is con- perature observation in Figure 40. With additional sistent for all cases with the exception of the case with EGR, the premixed peak temperature in the burned varying geometric compression ratios. While the to- zone reduces by a maximum of 377 K (LIVC 90) and tal trapped mass, inlet temperature and backpressure 216 K (EIVC 90). The other cases have a premixed were controlled, no appreciable change was found in peak burned zone temperature comparable or lower the intake pressure and exhaust temperature. Trends to baseline IVC. It is also interesting to observe that similar to Figure 24 from the base case are visible in the peak burned zone temperature through the diffu- Figure 36. sion combustion period reduces with Miller timing by the order of 70 K for LIVC and 35 K for EIVC. The Figure 37 shows the combustion parameter variations lower temperatures in addition to lower oxygen con- which were observed to have minor differences from centration (residual gas variation) explain the ISNO the base case. SOI varies marginally by advancing to reduction observations. maintain CA50 and the burn duration was observed to marginally increase. These could be attributed to the While the emission potential is understood, the slowing down of combustion due to additional EGR marginal loss in performance can be explained when and additional time required for mixing when com- observing the AHRR. In Figure 41 it can be seen that pared to the base case. Variations in ignition de- the peak of AHRR tends to phase later into the cycle lay and maximum rate of pressure rise were similar with Miller timing. and comparable to the base case. Maximum cylinder pressure and temperature showed small devia- This is consistent for both EIVC and LIVC. Although tions towards the IVC extremes. The change was to CA50 is controlled, having the AHRR peak later into the order of 3 bar and 20 K with extreme Miller (LIVC the cycle wherein combustion occurs in an increasing 90). volume through the expansion and leads to a loss in efficiency and performance [13]. Figure 38 provides more clarity on the factor of significance in this case, EGR. While it can be observed that the EGR rate was held constant (constant mass flow), a decrease in lambda and corresponding increase in % residual gas is observable. This can be explained by the overlap duration between the intake and exhaust valves which was constant through the evaluation (no change in opening of the intake valve or closing of the exhaust) and not accounted for in the control functions. Although the differential pressure across the cylinder is the same (controlled as scavenging pressure difference), the higher exhaust pressure with Miller as seen in Figure 36 would cause a higher momentary reversal of flow into the cylinder when the intake valve opens.

This explains a 4-5% relative increase in % residual gas content for a given EGR rate at the extreme IVC settings from baseline IVC. The impact of the fixed EGR rate is most prominent on the burned zone temperature and the ISNO. An ISNO reduction of 26% for EIVC 90 and 50% for LIVC 90 from base IVC (0 CAD) shows the significance of the Miller effect in this regard. The influence of the residual gas variation overestimates the isolated Miller effect on ISNO although the magnitudes are consistent with literature [16] on the NO reduction potential.

39 Figure 36: Observations for constant EGR rate:Set I

Figure 37: Observations for constant EGR rate:Set II

Figure 38: Observations for constant EGR rate:Set III

40 Figure 39: Observations for constant EGR rate:Set IV

Figure 40: Burned zone temperature with EIVC and LIVC

Figure 41: Apparent heat release rate with EIVC and LIVC

41 Performance Enhancement by Increasing Air improvement over the base case. Mass The performance potential highlighted in the base case as the PCP reserve with the Miller cy- The improvement does not appear to be significant cle was tapped by increasing the boost pressure and with the maximum benefits observed at LIVC 60 with hence the total trapped mass such that the resulting an ITE increase of 0.29% points over baseline (to PCP with Miller matches the base case PCP of 187.2 44.15%) which corresponds to an absolute IMEP in- bar. With the exception of the total trapped mass, all crease of 0.16 bar (to 25.01 bar) and an absolute other control functions were consistent as mentioned ISFC decrease of 1.23 g/kWh (to 189.61 g/kWh) from under Table 7. The injected fuel mass is kept constant baseline IVC. The reduced heat loss observed was at the baseline value of 279.8 mg/cycle. around 1.3% for EIVC 90 and 2.6% for LIVC 90 when compared to the base case while the swirl did not The observations are classified on a similar basis show an appreciable change. as performed previously and additionally includes the base case responses for parameters of interest to Figures 46 and 47 illustrate the reduced mass aver- contrast the effect of the treatment. The base case aged temperatures through the cycle which support responses are represented as dashed lines with the the observations of heat loss reduction with Miller tim- legend stating the parameter followed by ‘-base’. ing.

An increase in the boost pressure and corresponding When compared to the base case (Figures 30 and increase in the total trapped mass is visible in Figure 31) it can be observed that the peak temperature and 42. The total trapped mass increases by 3.5% with temperature through the expansion stroke are visibly a boost pressure increase to 5 bar for EIVC and by lower which also explains the lower exhaust temper- 7.5% with a boost pressure increase to 6.43 bar for ature to the end of the cycle. The peak temperature LIVC when compared to baseline. The motoring tem- drops by 48 K for EIVC 90 and 97 K for LIVC 90. perature remains similar to baseline while the motoring pressure curve flattens towards a higher motoring While the efficiency improvement due to reduced heat pressure with the added trapped mass. While the in- loss is understood, Figure 48 showcases the improve- let temperature is controlled, the exhaust temperature ment in combustion characteristics by observing the drops by 14◦C for extreme EIVC and 29◦C for extreme AHRR. It can be seen that with EIVC and LIVC, the LIVC. AHRR peak is phased earlier into the cycle which is beneficial for a given CA50. This benefit appears to A minimal SOI adjustment within 0.25 CAD is ob- be partially offset by the AHRR remaining marginally served in Figure 43 to maintain CA50 consistent. higher than baseline IVC further into the cycle (∼20 Also observable are a shortening burn duration and CAD). a marginal increase in ignition delay. The maximum rate of pressure rise decreases with Miller in this case The increase in lambda is the significant change due to better ignition conditions. The performance in- in this case with improvement of the ignition condi- crease can be visualised by the increase in PCP to tions due to additional trapped mass. The increase the target value while the peak cylinder temperature in oxygen concentration along with more acceptable shows a small drop with Miller timing to the order of 28 pressure and temperatures could explain the AHRR K and 57 K for extreme EIVC and LIVC respectively. trends observed.

Increasing the air mass along with the Miller effects’ reducing EGR dependence causes a noticeable increase in lambda by nearly 3% for EIVC 90 and 7% for LIVC 90 as can be seen in Figure 44. The corresponding residual gas reduction observed was by 0.5% points for EIVC 90 and 1% point for LIVC 90. ISNO appears to be controlled and the peak burned zone temperature is fairly constant.

The performance increase potential is largely inﬂu- enced by two factors in this case. Firstly, increasing the air mass ﬂow causes an increase in the pumping loss (PMEP) which is detrimental to ITE and hence IMEP and ISFC. Figure 45 depicts the PMEP increases from the base case by 3.3% and 7.5% for the extreme EIVC and LIVC timings. The increase in lambda with reduced burn duration and in-cylinder temperature implying lower heat losses improves the ITE and compensates for the increased PMEP. Over- all, the IMEP and ISFC curves for this case show an

42 Figure 42: Observations for performance increase with increased air mass:Set I

Figure 43: Observations for performance increase with increased air mass:Set II

Figure 44: Observations for performance increase with increased air mass:Set III

43 Figure 45: Observations for performance increase with increased air mass:Set IV

Figure 46: In-cylinder temperature with LIVC

44 Figure 47: In-cylinder temperature with EIVC

Figure 48: AHRR with LIVC and EIVC

45 Performance Enhancement by Increasing Fuel trapped mass constant, the relative increase of in- Mass The next case was to meet the base case jected fuel mass is larger and causes lambda to drop. PCP by increasing the injected fuel mass and to observe the potential performance improvement and its The impact on performance is illustrated in Figure 52. implications on other parameters. This case was An IMEP increase from baseline IVC is observable quantified in the reference study [16] to have provided for both EIVC and LIVC with a maximum increase a 5% IMEP improvement for EIVC (-70 CAD) and a of 3.8% (to 25.8 bar) for EIVC 90 and by 5.3% (to 10% IMEP improvement for LIVC (+80 CAD). 26.16 bar) for LIVC 90. Since air requirements are consistent with the base case, the PMEP remains un- Most of the air system parameters appear to be simi- changed. lar to the base case as can be seen in Figure 49. In- jected fuel mass increased by 4.2% for EIVC 90 and While a noticeable performance increase was ob- by 8.5% for LIVC 90 to meet baseline PCP. The in- tained, the ITE appears to have an more pronounced creased fuel mass caused the exhaust temperature drop from baseline (0 IVC) than the base case. This to raise by 19◦C for EIVC 90 and by 38◦C for LIVC drop in ITE leads to an increase in the ISFC. The 90. ISFC for the points with maximum performance increase was observed to be 191.66 g/kWh for EIVC A major influencing parameter in this particular case 90 and 196.68 g/kWh for LIVC 90. These correspond would be the injection profile. It must be noted that the to an increase in ISFC from baseline by 0.42% and injection profile is fixed and adjustments for engine 3% respectively. speed or change in injected fuel mass is only through variation of the injection duration (refer Figures 6 and The drop in ITE can be attributed primarily to the in- 7). creased heat loss due to higher in-cylinder temperatures and possibly the peak of heat release and in- Figure 50 depicts the change in SOI, EOI and burn creased burn duration due to the additional fuel mass. duration with the increase in injected fuel mass. SOI The heat loss shows an increase of 4.6% for EIVC 90 advances to maintain CA50 while EOI increases from and 8.8% for LIVC 90 from respective points in the the base case due to increase in injected fuel mass. base case. The increased temperature also leads to The burn duration shows an inverse trend to the base a worsening of working fluid property [11]. case with a maximum increase of around 2 CAD. PCP increases to the target value from the lower base case Figures 53 and 54 show the mass averaged in- values while the maximum cylinder temperature un- cylinder temperatures with both LIVC and EIVC to be derstandably increases from the base case due to the greater than the baseline. The higher temperature is additional fuel. The temperature increase is observed observed not only at the peak but also through the to be of the order of 30 K for EIVC 90 and 60 K for expansion stroke up to the opening of the exhaust LIVC 90. valve. The higher temperature through the expansion explains the higher heat loss and also the higher tem- The good ignition conditions at this point leads to perature at the exhaust. a fairly consistent ignition delay and maximum rate of pressure rise. The LIVC 90 case with the max- The AHRR in Figure 55 shows that the peak of heat imum increase in injected fuel mass shows a maxi- release advances earlier into the cycle for both the mum higher rate of pressure rise. LIVC and EIVC cases. This would mean that the efficiency should improve but it is also observable that While the air and EGR mass flow rates are similar to the heat release is higher beyond ∼18 CAD and later the base case, lambda shows an inverted variation into the cycle. The higher heat release later into the with Miller and increased fuel mass as can be seen in cycle could be attributed to the later EOI and greater Figure 51. A decreasing EGR dependence with Miller injected mass implying longer burn duration. was previously understood due to the lower process temperatures but in this case, a further reduction from the base case is visible.

This can be explained by the increase in fuel and hence the equivalence ratio (lower lambda) leading to a relatively richer mixture than the base case. As a richer mixture by deﬁnition does not contribute to NO formation, a lower EGR dependence than the base case is understandable for the same ISNO. Addition- ally, since the IMEP is expected to increase, an increase in absolute NO would be allowable so long as the ISNO is consistent.

While air mass does increase to keep the total

46 Figure 49: Observations for performance increase with increased injected fuel mass:Set I

Figure 50: Observations for performance increase with increased injected fuel mass:Set II

Figure 51: Observations for performance increase with increased injected fuel mass:Set III

47 Figure 52: Observations for performance increase with increased injected fuel mass:Set IV

Figure 53: In-cylinder temperature with LIVC

48 Figure 54: In-cylinder temperature with EIVC

Figure 55: AHRR with LIVC and EIVC

49 Performance Enhancement by Increasing Com- The Miller Cycle on a Serial Engine Con- pression Ratio The final performance increase figuration with T2S Miller evaluation on the trial on the single-cylinder configuration was by full engine was primarily performed to understand the means of increasing the geometric compression ratio similarities and differences involved when contrasting (GCR) in order to match the baseline PCP. The moti- with the single-cylinder approach and to quantify the vation behind this approach is to potentially increase charging system demands. While the single-cylinder the over-expansion effect of the Miller cycle by provid- analysis considered the isolated Miller effect for a ing an even higher expansion to effective compres- constant ISNO and EGR rate and also had 3 perfor- sion ratio. All other control parameters were main- mance enhancement cases, the full engine analysis tained to the target values specified in Table 7. was limited to 2 such sub-cases.

The major change in this case is illustrated in Figure The first case which forms the basis was the isolated 56 wherein the change in the GCR in order to meet Miller effect with constant ISNO. When considering the baseline PCP is shown. The GCR increases from the performance enhancement cases on the single- the baseline value of 17.3 to 17.72 for EIVC 90 and cylinder, it is clear that increasing the boost pressure 18.18 for LIVC 90. The DCR varies marginally from to meet PCP leads to extremely high intake pres- baseline due to the change in GCR. This change im- sures exceeding 6 bar while the case with varying ge- pacts the motoring pressure which flattens towards ometric compression ratio is not readily realisable in the baseline with an insignificant impact on the mo- practice. It was also understood that the maximum toring temperature. Inlet and exhaust pressure and performance enhancement was achieved with the in- temperatures were comparable to the base case and creased injected fuel mass case. Under these consid- the trapped mass was held constant. erations, the performance enhancement on the full- engine was only performed by increasing the injected The combustion parameters are largely comparable fuel mass to meet target PCP. to the base case with the exception of the PCP which reaches the baseline target as seen in Figure 57. Im- The results for the full engine evaluation with the T2S proving ignition conditions causes the maximum rate is split in two parts similar to the single-cylinder analy- of pressure rise to flatten towards the baseline value. sis. The first part will showcase and discuss the gen- eral observations and trends. Parameters that are No significant changes in the air and EGR flow rates in-cylinder in nature correspond to cylinder-1 of the were observed as seen in Figure 58. ISNO appears full engine while other parameters that would repre- to be controlled and the maximum burned zone tem- sent conditions for the full engine correspond to the perature is consistent. cylinder averaged values (where applicable). The assumption at this point is that cylinder-cylinder varia- Although the air and fuel parameters were not altered tions are insignificant and a separate section follows in this case, the change in GCR does have an im- the results and discusses the validity of this assump- pact on the performance parameters primarily as an tion. Figures from the corresponding single-cylinder improvement in the ITE which leads to a correspond- case will be mentioned at the end of an observation ing improvement in IMEP and ISFC as seen in Figure description as a reference for the reader. 59. The ITE shows a maximum increase of 0.24% points (to 44.1%) at LIVC 60 from the baseline ITE of 43.86% which corresponds to an absolute ISFC de- Miller Effect for constant ISNO Differing trends crease of 1.05 g/kWh (to 189.8 g/kWh) and an IMEP from the single-cylinder analysis are visible in Figure increase by 0.5% (to 24.98 bar). 61, 62, 63, 64 and 65. Additional observations for the serial engine analysis includes quantifying the T2S A heat loss reduction benefit is also observed and is systems demands and understanding the reduction to the order of 1.36% for EIVC 90 and 2.5% for LIVC in exhaust temperatures at EAT/muffler inlet. 90. This benefit is better understood when analysing the AHRR for the Miller cases as shown in Figure 60. In Figure 61, the motoring pressure and temperature show similar trends with the Miller cycle when The improvement in ignition conditions with the compared to the single-cylinder analysis although increase in GCR and the adjustment of SOI to inject the change is of equal magnitude for pressure while earlier causes the AHRR peak to occur earlier in the greater for temperature on both LIVC and EIVC ex- cycle and reduces the burn duration. This improves tremes. The trapped mass (cylinder averaged) ap- the efficiency on accounts of peak heat release pears consistent with the exception of LIVC 90 which phasing and lower heat release and hence in-cylinder shows a deviation within 1.5% as mentioned earlier. temperatures later into the cycle implying lower heat losses.

50 Figure 56: Observations for performance increase with increased geometric compression ratio:Set I

Figure 57: Observations for performance increase with increased geometric compression ratio:Set II

Figure 58: Observations for performance increase with increased geometric compression ratio:Set III

51 Figure 59: Observations for performance increase with increased geometric compression ratio:Set IV

Figure 60: AHRR with EIVC and LIVC

52 The reduced boost pressure demand for LIVC 90 (5.6 reduction is found to be around the same proportion bar compared to 6 bar for single-cylinder) is partly due as in the single-cylinder case and hence the improve- to the inability of the T2S to meet target trapped mass ments observed in performance parameters can be while EIVC 90 demands 4.8 bar similar to the single- attributed solely to the PMEP increase. cylinder analysis. A positive scavenging pressure differential is observed beyond EIVC 30 and for all LIVC The cooler heat rejection demanded for meeting con- timings. The backpressure does not exceed 3.75 bar troller targets on inlet temperatures are visible in the in contrast to over 6 bar in the single-cylinder case bottom right corner of Figure 64. The charge air due to imposed scavenging pressure difference (Re- cooler (CAC) has a near constant demand of 48 kW fer Figure 24 for equivalent single-cylinder case). while the EGR and interstage coolers (ISC) have a varying demand based on the EGR rate and LP stage The engine inlet temperature appears to be in con- boost. With Miller timing the EGR rate drops and trol while the exhaust temperature appears consistent the additional boost is provided by the LP stage and around 585◦C. The EAT inlet temperature (muffler) hence the EGR cooler demand drops while the ISC shows a reduction from baseline as expected. The demand increases. The negative sign indicates en- dashed line indicates the EAT inlet temperature from ergy lost from the frame of reference of the fluid . the base engine model (434◦C). The maximum reduction in EAT inlet temperature is to the order of The base engine model at the evaluation point had a 24◦C for EIVC 90 and 40◦C for LIVC 90 although it is heat rejection of 43 kW for the EGR cooler and 45 kW consistently above 380◦Cand may not be considered for the CAC. LP EGR with T2S reduces EGR cooler to have a significant impact on the SCR conversion heat rejection to less than 25kW (around 15 kW for efficiency. extreme Miller) while the ISC has a maximum heat rejection of 31 kW (LIVC 90). The total heat rejection Figure 62 depicts the injection parameters of cylinder- shows an increase between the base engine with a 1 of the full engine along with the combustion parame- one stage VGT and HP EGR (model) and the same ters. CA50 appears to be controlled with a marginally engine with T2S and LP EGR given the similar cooler higher SOI variation of 0.3 CAD from baseline when exit temperatures. The increase in heat rejection ca- compared to the single-cylinder case. The pressure pacity is around a maximum of 7% (6 kW) at LIVC reserve potential appears to be lower in the serial 90. engine with a maximum PCP margin of 5 bar (LIVC 90) while the peak temperature reductions are com- Figure 65 consolidates the compressor and turbine parable to the single-cylinder case. The ignition de- performance required for the T2S system. The con- lay, maximum rate of pressure, burn duration and EOI trol logic ensures a constant TPR for the HP stage show similar trends to the single-cylinder case (Refer and hence its demand power is constant with an ex- Figure 25 for equivalent single-cylinder case). ception at LIVC 90 accounting for the lower mass flow at that point. The HP turbine shows a reducing TPR Figure 63 illustrates the air and EGR variations in (inlet/exit) and wastegated flow as a result of the LP cylinder-1 of the serial engine with Miller timing. ISNO stage turbine increasing its TPR with Miller timing to appears to be controlled at the target with a fairly con- meet the additional boost (shown by increasing LP stant maximum burned zone temperature. Although compressor power and TPR). the trend of lambda and %residual gas appears similar, there is a marginal increase in lambda and de- The turbine sizing (orifice diameters) appears suc- crease in %residual gas for cylinder-1 when com- cessful for all points except LIVC 90 wherein the de- pared to the single-cylinder case (Refer Figure 26 for mand boost to meet the trapped mass condition is equivalent single-cylinder case). not met even when there is zero wastegated flow. In- creasing the HP stage TPR for LIVC 90 alone by re- The engine performance parameters are illustrated in ducing wastegated flow could resolve this under the Figure 64 along with the cooler heat rejection needs. defined turbine geometries. A significant change in trends can be observed with an increase in net IMEP by 5% ( by 1.26 bar) for EIVC 90 and 5.5% (by 1.39 bar) for LIVC 90. The IMEP increase can be explained by the PMEP becoming in- creasingly positive with Miller timing (+0.42 bar EIVC 90 and +1.55 bar LIVC 90) and a noticeable increase in ITE (Refer Figure 27 for equivalent single-cylinder case).

The ITE increase can be attributed to reduced heat loss and gas exchange loss which correspondingly improves the ISFC. The ITE increase by 2% points to over 46% for extreme Miller cases which corresponds to nearly 10 g/kWh reduction in ISFC. The heat loss

53 Figure 61: Observations for full engine with T2S Miller effect at constant ISNO :Set I

Figure 62: Observations for full engine with T2S Miller effect at constant ISNO :Set II

Figure 63: Observations for full engine with T2S Miller effect at constant ISNO :Set III

54 Figure 64: Observations for full engine with T2S Miller effect at constant ISNO :Set IV

Figure 65: Observations for full engine with T2S Miller effect at constant ISNO :Set V

55 Miller and Cylinder-Cylinder Interaction This the compression stroke which becomes more signifi- section builds on the results obtained under the previ- cant with increasing Miller timing. The change in to- ous case (full engine with T2S at constant ISNO) with tal trapped mass naturally impacts the residual gas a focus on the potential impact that cylinder-cylinder content and lambda although the intensity of variation interactions could have on the variation of parame- appears more marked with lambda than with residual ters observed from cylinder-1. The assumption under gas content. the previous section was that the variation of cylinder dependent parameters would be insignificant among The lower significance of residual gas content varia- cylinders and that the trends observed for cylinder-1 tion could be attributed to a lower EGR dependence with Miller timing would hold true for the other cylin- with Miller and due to a LP EGR circuit which provides ders. a more uniform mix of burned gas in the fresh charge.

Upon investigation of the parameters on a per cylinder The impact of these variations on the performance pa- basis, it was found that the assumption holds only for rameters are illustrated in Figure 74. The dashed line a portion of the results and understandably so. The in the individual plots indicates the cylinder averaged EIVC cases and the baseline IVC case did not show values that were referenced under the case results. any significant variation in the cylinder dependent pa- While it is clear that variations in ITE would naturally rameters and hence the assumption of cylinder-1 be- impact the IMEP and ISFC variations, it is interesting representative of all cylinders holds here. ing to observe that the PMEP shows consistently low variations for all LIVC timings. Thus the performance On the other hand, with LIVC, a portion of the in- parameter variations can be attributed towards those gested charge of one cylinder is pushed back into the parameters that cause ITE variation and shows that intake manifold due to IVC late(r) into the compres- the first 3 cylinders perform better than the latter 3. sion (with more extreme IVC) stroke. This “push back” of charge from the cylinder causes an uneven distri- The pressures and temperatures that impact ignition bution in the total trapped mass, air (oxygen content) and those that are a result of combustion are illus- and residual gas (EGR) among the cylinders. The trated in Figure 75. While it is understood that Miller variation of these key parameters from the air han- timing would lead to worsening ignition conditions dling side can be termed as the “cause” for the vari- which is clear from the average shift of parameter ation in all other parameters (“effect”) that are depen- values with Miller, the variation in conditions among dent upon these fundamental parameters. cylinder appears significant for LIVC 90. The motoring pressure shows a decreasing trend among cylin- The cylinder-cylinder variation significance will be dis- ders for a given Miller timing (lower trapped mass) cussed for both the serial engine base case (constant while the motoring temperature is increasing proba- ISNO) and the performance enhancement case on bly due to the higher temperature of ingested charge the full engine. (Figure 73).

PCP and peak temperature variations are also high- Cylinder-Cylinder Variation: Serial Engine Base lighted in Figure 80 with the average PCP showing a Case Figures 73, 74, 75, 76 and 77 illustrate the decline with Miller timing and an increase in variations variation of the cause and effect parameters on a to the order of 10 bar for LIVC 90. Peak tempera- cylinder-cylinder basis. The Figures are placed under ture variations appear to be more pronounced in the the appendix to maintain image resolution and would cylinders with a relatively richer mixture with temper- aid in understanding the qualitative arguments on the ature variations tot he order of 150◦C. The tempera- signiﬁcance of cylinder-cylinder variations on engine ture variations would naturally affect the cylinder heat performance. transfer and hence the cylinder efﬁciency.

In Figure 73, the variation of total trapped mass The SOI variation to maintain CA50 between cylin- among the cylinders is evident with increasing Miller ders and different Miller timings is illustrated in Figure timing (LIVC 30 to 90). The dashed line indicates the 76. While CA50 appears to be well maintained, it is control target which relies on cylinder averaged feed- clear that an individual control on SOI was necessary back. It can be seen that with increasing Miller timing, to achieve this. Initial trials in the study had cylinder the cylinder filling of the first 3 cylinders is typically average based control for CA50 and SOI which lead better than the latter 3 and that while the mean value to PCP variations of the order of 15 bar and a drop of of the readings are consistent the standard deviation PCP to as low as 130 bar for LIVC 90. The individual is increasing. SOI control is considered crucial and has a significant impact when considering Miller timing on a serial en- The filling anomaly can be partly explained by the gine setup. increase in temperature at points in the manifold directly leading to the respective cylinder ports. The Figure 76 also illustrates the variation of ignition delay higher temperature could be from charge that enters and burn duration. Ignition delay appears to be rela- a cylinder and is pumped back to the manifold through tively less sensitive to the variations probably due to

56 the operating point (1000RPM full load) but still shows impact on SCR conversion efficiency. a marginal increase in magnitude and variation with Miller timing. The cylinders with a higher charge tem- In Figure 67, SOI advances due to the increase in perature and relatively richer mixture (cylinders 5 and injected fuel mass along with a marginal increase in 6) appear to have a lower ignition delay and a longer burn duration with Miller timing. The cylinder aver- burn duration. aged PCP is depicted and appears to meet the target value for all cases. The maximum cylinder(1) temper- Figure 77 sums up the cylinder-cylinder variation ature which is expected to increase with Miller tim- analysis for the constant ISNO case of the serial en- ing due to additional fuel shows a decrease for LIVC gine. The expected higher heat losses from cylinders cases. This will be further investigated in the cylinder- with higher temperatures appears to hold true. This is cylinder variations section. Ignition delay and maxi- supported by the swirl variation which shows higher mum rate of pressure rise show comparable trends to values for cylinders 4,5 and 6. The higher swirl could the serial engine base case. explain better mixing and hence lower ignition delay while also increasing the heat transfer [13]. The air and EGR metrics are depicted in Figure 68 wherein the ISNO appears to be under control with The variation in ISNO between cylinders is perceiv- a similar observation of near consistent maximum able as cylinders 5 and 6 typically have higher values burned zone temperature. The air and EGR mass than target while cylinders 1-4, particularly cylinders flow rates are also comparable to the serial base en- 1 and 2 have lower values than target. Since cylinder gine case. While the residual gas variation (cylinder- averaged ISNO is the controller feedback this varia- 1) with Miller timing appears consistent with previous tion in ISNO is not visible from the serial engine per- observations, lambda shows a new trend of being spective nor does it get corrected by the variation in nearly constant through EIVC and baseline cases and combustion and performance parameters. increasing for the LIVC cases. This behaviour could a result of cylinder interactions and will be analysed in the cylinder-cylinder variations section. Performance Enhancement by Increasing Fuel Mass Similar to the single-cylinder analysis of per- The performance enhancement potential of the serial formance enhancement by increasing the injected engine with T2S and Miller timing using increased fuel fuel mass, the full engine analysis was also based injected mass is illustrated in Figure 69. IMEP shows on PCP feedback to the the injector controller. The a significant increase of ∼14% (to 28.93 bar) for LIVC PCP feedback was cylinder averaged and was a sin- 90 and ∼10% (to 27.89 bar) for EIVC 90 while the gle signal to all cylinders. The control was not im- PMEP variation is consistent with serial engine base plemented on a cylinder basis to avoid compounding case. Within the observed increase, the additional the combustion variability due to inherent variations fuel mass contributes to a 5.5% IMEP increase for in the oxygen content and EGR between cylinders as EIVC 90 and nearly 9% increase for LIVC 90. The was understood from the previous section. The Miller (positive) PMEP increase accounts for the remaining effect at constant ISNO on the full engine will be re- benefit which was observed in the serial engine base ferred to as the “serial engine base case” to contrast case. the performance enhancement effects on the full engine with T2S. ITE and ISFC show similar tends to the serial engine base case with ISFC still showing a net reduction. Figure 66 shows the increase in injected fuel mass for ISFC increases by 0.6% for both extremes of IVC with the performance enhancement (PCP increase). As the base case. The proportion of heat loss appears to no changes were made to the air system demands, have increased (7-10% at extreme IVC) due to higher these attributes show comparable trends to the full in-cylinder temperatures. Air system cooler heat re- engine base case. Injected fuel mass shows an injection for the EGR cooler, ISC and CAC are consis- crease by 6% for EIVC 90 and 8% for LIVC 90 from tent with the serial engine base case observations. baseline (0 IVC). An interesting observation here is that the trapped mass target appears to be in control The T2S performance illustrated in Figure 70 is com- for LIVC 90 and the T2S achieves this due to the in- parable in terms of compressor performance while crease in exhaust enthalpy (evident through tempera- the turbine characteristics show improvement due to ture increase ∼25◦C). higher enthalpy exhaust gas in this case. More flow is wastegated by the turbines or in other words the The positive scavenging pressure differential is main- compressor demands are met with relative ease when tained and the inlet temperature is visibly under con- compared to the serial engine base case. trol. The EAT inlet temperature shows a similar trend to the serial engine base case but has an offset to- Figure 70 also shows that the current sizing of the wards a higher temperature due to the additional fuel T2S is sufficient to meet the trapped mass demands and exhaust energy. The LIVC 90 case shows a even for the LIVC 90 case which it was unable to do lower EAT inlet temperature from baseline by 12◦C so without the additional exhaust enthalpy. at 422◦C which is not expected to have a significant

57 Figure 66: Observations for full engine with T2S Miller effect and fuel mass based performance increase :Set I

Figure 67: Observations for full engine with T2S Miller effect and fuel mass based performance increase :Set II

Figure 68: Observations for full engine with T2S Miller effect and fuel mass based performance increase :Set III

58 Figure 69: Observations for full engine with T2S Miller effect and fuel mass based performance increase :Set IV

Figure 70: Observations for full engine with T2S Miller effect and fuel mass based performance increase :Set V

59 Cylinder-Cylinder Variation: Serial Engine Perfor- rial engine base case (Figure 76). The variation re- mance Case The cylinder-cylinder variation effects mains comparable and ensures CA50 is maintained for the serial engine under the performance enhance- to target. The ignition delay and burn duration show a ment cases are expected to follow trends from the marginal increase with the variation trends remaining serial engine base case (constant ISNO) discussed comparable. previously. The major variable here is the increase in injected fuel mass which is expected to affect the The swirl and heat transfer variations in Figure 82 are outcomes of combustion. Better performance of the comparable to the serial engine base case (Figure T2S implies the air system variables would vary par- 77) with the magnitude of heat transfer showing an ticularly for the LIVC 90 case. The observations for increase. The trend with ISNO remains comparable depicted through Figures 78, 79, 80, 81 and 82 which to the serial engine base case with the magnitude of are also placed under the appendix for reader refer- ISNO showing a decrease for the LIVC 90 case alone. ence. While the trend for maximum rate of pressure rise In Figure 78, it can be seen that while the total trapped is conserved across the cases, the performance en- mass variation trends are consistent with the serial hancement case shows an increase in magnitude for engine base case (Figure 73), the total trapped mass the LIVC 75 and LIVC 90 cases. shows an increase for all the cylinders under LIVC 90 while the observations remain comparable for other CONCLUSIONS Miller timings. While inlet temperature and residual gas content variations are also comparable, the A predictive combustion model (DI Pulse) was incor- lambda variations retain the trends of the serial en- porated into the existing base engine model to enable gine base case and show a consistent drop in magni- studies with varying combustion characteristics. The tude due to the increase in injected fuel mass to meet model was calibrated and validated with the limited PCP. dataset available and through sourcing of relevant information from literature. The predictive capability ap- The performance parameter variation shown in Figure peared reasonable in accordance to studies that fo- 79 retains the trends shown for the serial engine base cused on this aspect although deviations in burn rate case (Figure 74) while the magnitude of ITE shows a amplitude and phasing were evident. An attempt at drop, the magnitude of ISFC and net IMEP show an understanding the predictive burn rate phasing was increase. This also implies that the magnitude of vari- made with respect to the IDM but the trials did not ation shows an increase (ﬁgures are scaled to appear lead to conclusive results. similar). In 80, the motoring pressure and temperature variations are comparable to the serial engine While the performance predictions of combustion base case (75) with the LIVC 90 case showing rela- were based on calibration and were validated, the tively higher pressures and temperatures owing to the emission models lacked the necessary data to base additional boost (trapped mass target met). the calibration upon. This led to incorporating only a NO emission model based on the extended Zeldovich While the PCP between cylinders should ideally be mechanism with default calibration multipliers. A cal- around the target (187.2 bar), the cylinders with a ibration trial was made for NO with data from litera- higher trapped mass (1,2 and 3) typically show higher ture to understand the nature of deviation between pressures (greater than 190 bar for LIVC 75 and 90) calibrated and uncalibrated cases. The understand- while the remaining cylinders tend to a lower PCP (to- ing from the trial was that the uncalibrated NO model wards 180 bar for LIVC 75 and 90). This variation (default multipliers) overestimated NO by a near con- was not addressed through individual PCP control for stant gain for a given operating condition. The trials the cylinders to avoid further increasing the combus- did not involve EGR and hence its impact on the NO tion condition variability as the cylinders with lower model response was not ascertained. PCP are already relatively richer and increasing PCP would mean greater injected fuel mass which further Upon calibrating the DI Pulse combustion model for reduces lambda. Similarly, cylinders with higher PCP performance and deciding to use the uncalibrated ex- having a relatively higher lambda value wherein the tended Zeldovich mechanism for NO, a virtual test rig injected fuel mass would have to be reduced to match was built for the single-cylinder and the serial engine PCP target would further increase lambda. with T2S conﬁgurations of the base engine model.

The variation in peak cylinder temperature is similar The single-cylinder model had controller implementa- to the serial engine base case while the magnitude tion through simple SISO representations of the total for the performance enhancement case shows an in- trapped mass, inlet temperature, ISNO, CA50, scav- crease due to the additional fuel mass. enging differential pressure and the PCP. The model was tuned to match the base engine model at the The injection parameter variation depicted in Figure evaluation point with no Miller timing showing devi- 81 shows an increased advance in SOI due to in- ations within 1% for the air and EGR mass ﬂow, increased injected fuel mass when compared to the se-

60 take and exhaust pressures and the net IMEP. The 5.3% (to 26.16 bar) for LIVC 90. This was lower when PMEP match required additional tuning and the devi- compared to the literary reference which showed a ation was reduced to less than 1.5%. 5% and 10% improvement potential for the respective extremes in Miller timing [16]. 5 cases were evaluated under the single-cylinder Miller analysis with the constant ISNO case as the While the LIVC 90 case shows the maximum per- base case. The major observation from the constant formance increase potential from baseline, it comes ISNO case was the PCP reserve available for possi- at the cost of a lower ITE of 42.56% (1.3% point ble performance enhancement with Miller timing. The decrease) and higher ISFC of 196.68 g/kWh (5.83 maximum PCP reserve for EIVC (90 CAD) was 4 bar g/kWh increase). The most efficient point under this and for LIVC (90 CAD) was 8 bar which were consis- case is at EIVC 60 with an ITE of 44.01% (0.15% point tent but lower than observations in literary references increase), an ISFC of 190.21 g/kWh (0.64 g/kWh de- (maximum 10 bar [16]). crease) but a lower increase in IMEP from baseline to 25.13 bar (0.29 bar/1.16% increase). A distinct case on the NO reduction potential of Miller timing was investigated at a constant EGR rate of The increase in GCR to meet the PCP target lead 17.58% which was obtained from the baseline IVC to an increase from the base engine GCR of 17.3 EGR rate settings. The NO reduction potential ap- to 17.72 for EIVC 90 and 18.18 for LIVC 90. In this peared significant with a maximum reduction in ISNO case the maximum efficiency and performance point by 26% for EIVC 90 and 50% for LIVC 90 from the coincide at LIVC 60 with improvements from baseline baseline (IVC 0 CAD). The results are not purely due amounting to a net IMEP increase of 0.5% (to 24.98 to Miller timing due to varying residual gas content as bar), an ITE of 44.1% (0.24% point increase) and an a result of the valve overlap. This effect was not coun- ISFC of 189.8 g/kWh (1.05 g/kWh decrease). tered by incorporation into the control functions but the observations appear consistent although higher The serial engine model was built with a simplified than with literature which showed a 15% reduction on representation of a T2S system through the use of EIVC 70 and 34% reduction for LIVC 80 although for isentropic efficiency based compressor and turbine BSNOx [16]. While the NO reduction potential was models. The T2S sizing and serial engine model tun- significant for extreme Miller timing, the Miller loss being achieved similar air and EGR flow rates within 1% comes prominent and causes a drop in ITE and hence of the base engine model while the net IMEP was the IMEP with a corresponding increase in ISFC. within a 1.5% deviation margin at the evaluation point. The PMEP decreased by 44% from baseline owing to Performance increase to match the PCP reserve ac- backpressure originating purely from the T2S (no HP counted for the 3 remaining single-cylinder cases. EGR) while the intake and exhaust pressures reduce The approaches to match the target PCP was through within 5 and 10% from the base engine values. increased trapped mass (boost pressure), increased fuel injected mass and increasing the GCR. Indepen- The serial engine required additional controller devel- dent SISO controllers were implemented to meet the opment for wastegating the HP and LP turbines to PCP target across the cases. meet the TPR target of the HP stage and the trapped mass target with the LP stage. Inlet temperatures of The increase in total trapped mass led to boost pres- the individual compressors was controlled in addition sures over 6 bar for LIVC 90 and consistently ex- to the engine inlet temperature. The CA50 control ceeds the base case boost pressure for Miller tim- through SOI was performed on an individual cylinder ing ±60 CAD and above. LIVC 60 gives the best basis to reduce combustion variability and maintain performance benefit from baseline IVC with an ITE reasonable stability which appeared to be necessary of 44.15% (0.29% point increase) an ISFC of 189.61 through the model tuning trials. g/kWh (1.23 g/kwh decrease) and a net IMEP of 25.01 bar (0.16 bar increase). The PMEP at this The serial engine evaluation with Miller involved only point was higher than baseline at -0.426 bar which 2 cases in contrast to 5 in the single-cylinder analysis. was marginally higher than the base case due to the The serial engine base case was set similar to that of additional mass flow. the single-cylinder analysis with ISNO constant and Miller timing as the sweep parameter. For the per- The increase in injected fuel mass was proportionate formance increase trial with Miller on the serial en- to the PCP reserve from the base case with Miller tim- gine, only the increased fuel mass case was analysed ing which lead to an increase by 4.2% for EIVC 90 and owing to the very high boost pressures (over 6 bar) 8.5% for LIVC 90 from the baseline value to meet the observed under the single-cylinder analysis and the target PCP. A consequence of increased fuel mass practical limitations in realising a variable GCR along was an increase in the exhaust temperature by 19◦C with the observation that the increased fuel mass for EIVC 90 and 38◦C for LIVC 90 which would aid the case showed the maximum performance increase po- turbocharging and EAT systems. The maximum per- tential under the same constraints. formance increase witnessed from baseline was an IMEP increase by 3.8% (to 25.8 bar) for EIVC 90 and A significant observation under the serial engine trial

61 was that a positive scavenging pressure differential A follow-up study was performed to check the compa- developed with Miller timing. This was owing to the rability of Miller evaluation on the single-cylinder and absence of a HP EGR circuit and a high efficiency serial engine models by applying the backpressure T2S system. This inversion of the PMEP trend for the boundary condition of the serial engine performance serial engine case biases the results observed favour- increase case for the single-cylinder model controller ing the serial configuration when compared to results target. The comparison is presented under the ap- from the single-cylinder analysis. pendix in Figure 72 and shows the single-cylinder model predictions to be representative of the cylinder For the constant ISNO (serial engine base case), averaged values of the serial engine with deviations the boost pressure dropped when compared to the within 1%. single-cylinder equivalent with maximum pressures of 4.8 bar for EIVC 90 and 5.6 bar for LIVC 90. The The heat transfer rate and swirl appear to show a trapped mass target at LIVC 90 fell short by 1.5% ow- larger deviation of around 3% but it must be noted ing to the T2S systems’ inability to meet the required that the cylinder dependant values depicted in Fig- performance even at near zero wastegate condition ure 72 are from cylinder-1 of the serial engine and of the turbines. The EAT inlet temperature showed the cylinder-cylinder interaction analysis (Figure 82) a maximum drop of 50◦C from a baseline value of showed the swirl and heat transfer rate of this cylin- 434◦C and was consistently above 380◦C which im- der to be on the lower side. It can be concluded that plies that SCR conversion efficiency will not be signif- when serial engine relevant boundary conditions are icantly impacted at this point. used on a single-cylinder equivalent, the response of the single-cylinder correlates well with the cylinder av- Although this was considered a base case, a perfor- eraged serial engine response. mance improvement was observed due to the positive PMEP with Miller. A comparable net IMEP increase An investigation into the significance of cylinder- of 5-5.5% (to 26.5 bar) was observed for the extreme cylinder interactions due to Miller timing was carried Miller cases with the LIVC extreme showing the max- out. Observations on a cylinder basis showed that the imum increase from baseline. The PMEP turned pos- EIVC cases, baseline IVC and mild LIVC (30 CAD) itive and showed an increase of 0.6 bar (to 0.42 bar) cases did not show any significant variation in param- for EIVC 90 and an increase of 1.73 bar (to 1.55 bar) eters between cylinders. from baseline. Both EIVC and LIVC 90 have com- parably high ITE but LIVC 90 shows maximum ITE in- With greater LIVC timing (>30 CAD), the push back crease from baseline at 46.9% (2.45% point increase) of charge from a cylinder to the intake manifold and an ISFC of 178.5 g/kWh (9.81 g/kWh decrease). skews the distribution of the trapped mass, oxygen and residual gas content and charge temperature be- The IMEP and ITE at the baseline for the serial engine tween cylinders. The variation in these parameters base case show higher values than the single-cylinder is cited as the root cause for the variations observed equivalent but the positive PMEP increase with Miller in other parameters such as performance and emis- appears to amplify the benefits. The increased boost sions. It is clear that with increasing LIVC, the mag- demand with Miller timing achieved through the T2S nitude of the variations increased and was observed lead to the total heat rejection in the coolers (EGR, for the serial engine performance increase case as ISC and CAC) to increase by a maximum margin of well (with higher magnitude of variation). The PMEP 7% (6kW) from a baseline demand of 88 kW (EGR was the only parameter that did not show appreciable and CAC from base model) at LIVC 90. variation between cylinders with late Miller timing thus directing the observed variations to be a direct result The serial engine performance enhancement case of the combustion variability among cylinders. showed a fuel mass increase of 6% and 8% for EIVC 90 and LIVC 90 respectively. The trapped mass tar- DISCUSSION AND FUTURE WORK get at LIVC 90 that was not met in the previous case was now achieved owing to the additional exhaust There are several aspects concerning modelling and gas enthalpy with the increase in injected fuel mass. evaluation considerations through the study that limit The maximum boost pressure increased to 5.72 bar and influence the observed results to a varying de- to meet the trapped mass target at LIVC 90. EAT in- gree and must be accounted for to understand the let temperature falls short of baseline only at LIVC 90 ◦ basis of the results obtained. The factors considered only by 12 C and is hence inconsequential to the SCR crucial to enable this understanding will be described conversion efficiency. in this section along with possible actions that can be taken to improvise on the current study. A significant performance increase was witnessed at the extreme Miller points with the maximum IMEP at LIVC 90 increasing by nearly 14% over baseline to DI Pulse Calibration The study strongly relies 28.9 bar. This point also had the maximum ITE in- on a predictive combustion model and naturally the crease for the case with 46.6% (2.18% point increase) capability of the calibrated model is a factor that in- and the least ISFC at 179.67 g/kWh (8.76 g/kWh).

62 fluences the results derived from the combustion pro- sion models to perform insightful trade-off analy- cess which is central to a combustion engine. Factors sis of concern include the injection profile sourced from literature and its utilisation through linear scaling for varying speeds and loads (injected fuel mass). Base Engine Friction Characterisation The study considered performance and emission pa- The injection pressure for operating points used in rameters as indicated specific in place of brake spe- the calibration was not available and could only be cific. This was modified through the study once it was referenced with the injection pressure profile theoret- understood that the friction model of the base engine ically required to meet the imposed burn rate (non- was not of sufficient detail and due to the need to as- predictive) for the cases in the base model. While this sume the mechanical efficiency of the single-cylinder may appear to have worked in the current study, GT- engine to ensure comparable results between the Power has emphasised the need for case appropriate single-cylinder and serial engine cases. With the data injection profile information in the combustion model currently available, it was not possible to characterise calibration to obtain reliable outcomes [20]. The in- friction in the base engine model and hence to avoid jection profile limitation also extended to the non- errors in this aspect, the indicated specific terms were utilisation of points with pilot injections which would opted for. further inhibit the predictive capability of the model to be representative of the combustion behaviour of the To understand parameters on a brake specific scale, base engine. it would be beneficial to obtain relevant data and characterise engine friction through simplified models like In addition to the injection profile limitation, the DI the Chen-Flynn model which was only partially and Pulse model calibration could only be performed on inadequately enabled in the base model. a limited number of data points which was below the recommended guidelines. Furthermore, the default calibration process is performed as a single objec- Intake Valve Profiles for Miller Timing tive optimisation function with the target variable be- When considering the valve profiles used through the ing to reduce the burn rate error. While this appears study, the LIVC cases only required an increase in the to be an acceptable and verified means of ensuring peak dwell period to match the delayed IVC. In con- reasonable predictive combustion capability, studies trast, the EIVC profiles required scaling of the orig- such as [26] which focused on the DI Pulse calibra- inal intake valve profile to maintain the acceleration tion and predictive capability found that the calibra- consistent. Literature shows different ways of achiev- tion does not account for the deviations in PCP, CA at ing the EIVC either with the use of trapezoidal pro- PCP, CA10, CA50 and CA90. files from electric actuators [16] which maintain the peak lift and adjust valve opening angle or a similar While combustion calibration with respect to in- approach to what has been used in this study but cylinder pressure was possible with the limited num- with relatively higher peak lifts and hence accelera- ber of data points, predictions on emissions was tions [15]. largely limited by the absence of corresponding data for the base engine model. The extended Zeldovich The approach to achieving EIVC would affect the mechanism was used uncalibrated and since ISNO pumping work required to meet a certain trapped was a control parameter, the NO predictions would mass target and would also affect the nature of in- naturally affect the EGR needs and the performance cylinder flow. It must also be mentioned that the open- enhancement potential with Miller timing. HC, CO and ing of the intake valve and all characteristics of the PM emissions were not modelled to avoid further in- exhaust valve were kept consistent with baseline. consistencies with uncalibrated emission models. For a given EIVC timing, influence of different valve The recommended actions with regard to making the profiles could be evaluated to understand if the valve DI Pulse model robust would be acceleration limitation significantly increases pumping work to meet a certain trapped mass target. As- suming that calibrated emission models are available, • calibrate against minimum number of data points trade-off studies between NOx and PM for instance as recommended by GT (minimum 25 points) could be performed to understand valve profile influence on in-cylinder air flow (reduced order) and com- • utilise operating point appropriate injection pro- bustion. files for a more physically relevant model

• investigate the possibility to perform the calibra- Controller Considerations The controllers tion as a multi-objective function to overcome the used in the study functioned reasonably well but were drawbacks mentioned in [26] or report the com- not optimal in terms of runtime. Incorporating con- bustion model performance on these aspects trollers as SISO required greater modelling consider- • Calibrate the NO model and other relevant emis- ations to avoid instability and unnecessary controller

63 interaction. Having more advanced control schemes The lack of HP EGR in the serial engine model as MIMO and perhaps coupling with an external mod- favours the T2S system in its current configura- ule such as Simulink could enable shorter runtimes tion and does not impact EAT inlet temperature as and more efficient control. This would also enable severely as expected at the onset of this study. A better transient analysis of the cases. HP EGR circuit will deprive the T2S of some high enthalpy exhaust gas thereby making it harder for the The controller targets derived from the base engine turbines to meet a given boost demand. This may model are also a source of discrepancy. This is es- lead to turbine sizing to tend towards smaller ori- pecially true for control parameters that relied on the fice diameters which implies higher backpressure and DI Pulse model predictions in the base engine model. lower EAT inlet temperatures. Parameters such as the baseline ISNO and CA50 targets were set based on outcomes from the predic- While the modelling of both configurations (single- tive combustion being incorporated into the base encylinder and serial engine) with an LP EGR circuit gine model. While there were no means to ascertain for reducing controller complexity provided a basis for ISNO deviations from the true value, the CA50 target comparable models, the controller targets imposed showed a 0.8 CAD shift later into the cycle with the were derived from the base model with a HP EGR calibrated DI Pulse model when compared to the non- circuit. This implies that while the single-cylinder con- predictive (imposed burn rate) model originally incor- figuration had an imposed scavenging pressure dif- porated in the base engine model. Although the CA50 ferential (negative) from the base model, the serial target is kept consistent for all evaluation cases, the engine shows a positive scavenging pressure differ- later placement of CA50 would limit the efficiency and ential due to the backpressure originating purely from performance benefits that would arise when phased the turbines of the T2S. appropriately earlier into the cycle. In other words, the serial engine does not require a Controller feedback for the serial engine was more higher backpressure to meet the EGR demands re- challenging in deliberating between cylinder averaged quired to maintain the ISNO at target. This incon- signal feedback or individual controls for each cylin- sistency biases results to be more favourable for the der. The cylinder-cylinder variations were reduced in serial engine and underlines the need to incorporate magnitude by the use of individual controllers for com- the HP EGR circuit into the serial engine model for a bustion phasing with independent SOI adjustments fair comparison. Future investigations could consider while trapped mass, ISNO and injected fuel mass dual loop EGR with T2S to utilise the benefits of the were all cylinder average based controls. The result LP circuit which is favourable for steady-state oper- of this is visible variations in the PCP, cylinder ISNO ation while the HP circuit would help meet transient and trapped mass although it is the inconsistencies EGR demands. in the air handling system when LIVC is introduced which causes the variations. Cylinder Interaction Investigation The The use of control functions to address the variations analysis on the impact of interaction between cylin- could be investigated although it does not appear that ders producing a variation among them was quali- control functions can account for the root cause of the tatively established. The results provide a basis for variations which is mechanical in nature. considering cylinder interactions as a cause of concern for LIVC cases. The cylinder interaction issue is typically not featured in literature due to the lim- T2S Modelling and EGR Circuit Consider- ited number of Miller analysis on serial engines. In- ations The T2S implemented in the model is very dividual investigations could be performed to mitigate simplistic and the sizing of turbines does not have a the variation among cylinders either through intake “base” orifice diameter representation to understand manifold design, valve timing events or corrective con- turbine sizing in terms of orifice diameter. The in- trol functions on the combustion for LIVC cases that termediary step of characterising the base models’ show promise on the cylinder averaged steady-state single-stage VGT as an equivalent orifice would have results. As previously mentioned, the root cause of provided a basis to understand if the current HP and the variations are due to the non-uniformity in the to- LP stage sizing is realistic or biased towards steady- tal trapped mass, oxygen content and residual gas state operation. The stage efficiency and work-split content among cylinders and design considerations are also subject for further discussion but would per- to address this could be investigated for the LIVC tim- haps have more clarity if the T2S is built with clearly ing of interest. defined goals.

Characterising the single-stage VGT as an equivalent oriﬁce diameter could be a step towards better repre- senting and understanding physically relevant turbine side considerations for the T2S.

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66 APPENDIX

Figure 71: In-cylinder condition initialisation for CPOA and DIPulse calibration

A1 Figure 72: Single-cylinder and serial engine (-F) response comparison for performance increase by fuel mass

Figure 73: Serial engine constant ISNO cylinder-cylinder variation:dataset I

A2 Figure 74: Serial engine constant ISNO cylinder-cylinder variation:dataset II

Figure 75: Serial engine constant ISNO cylinder-cylinder variation:dataset III

A3 Figure 76: Serial engine constant ISNO cylinder-cylinder variation:dataset IV

Figure 77: Serial engine constant ISNO cylinder-cylinder variation:dataset V

A4 Figure 78: Serial engine performance increase cylinder-cylinder variation:dataset I

Figure 79: Serial engine performance increase cylinder-cylinder variation:dataset II

A5 Figure 80: Serial engine performance increase cylinder-cylinder variation:dataset III

Figure 81: Serial engine performance increase cylinder-cylinder variation:dataset IV

A6 Figure 82: Serial engine performance increase cylinder-cylinder variation:dataset V