her fortitude despite the compressed time schedule imposed upon her. Much ..... Figure 4.10: 3-D map of switching mode between 2-stage and single-stage charging ... 53 ..... 2000 RPM initial condition to full-load at constant engine speed. ..... Other common developments to both Diesel and Otto engines can be captioned.
Development and Transient Performance Simulation of a 2-Stage Turbocharged DI Otto Engine Concept
Master Thesis
David H. Oh 268616
Institute for Combustion Engines Aachen
Prof. Dr.- Ing. S. Pischinger RWTH Aachen University
Advisor: M.Sc. Sandra Glück
Submitted on: 22.12.2009
Declaration Name, First Name:
Oh, David
Matriculation Number: 268616
Title of Work:
Development and Transient Performance Simulation of a 2-Stage Turbocharged DI Otto Engine Concept
I hereby declare that I have composed the submitted work, including all attached materials, by myself. None other than the quoted tools and bibliography has been used; this applies for all source types.
This work in the same- or similar form has not yet been submitted.
I understand that breach of this declaration and deliberate misrepresentation may result in rejection of this work.
Location, Date
Signature
Acknowledgements The author wishes to express sincere gratitude to Sandra Glück for her advisement, review and constructive suggestions for the improvement of this work, especially for her fortitude despite the compressed time schedule imposed upon her. Much appreciation is also due to Christof Schernus for his guidance, insight and leadership from over two years of working together. Thank you, Professor Stefan Pischinger, for giving me the opportunity to have been part of FEV dating back to 2000 and the world-class institutions that are the VKA and RWTH Aachen University since 2005.
Dedicated to Mom and Dad, who have encouraged me to pursue my dreams to the greatest possible extent and who have prayerfully endured this journey with me through long periods of separation.
Contents
I
Contents 1
INTRODUCTION .................................................................................................... 1 1.1
1.2 2
Motivation ....................................................................................................1 1.1.1
Trend of Downsizing .............................................................................4
1.1.2
Downsizing Challenges ........................................................................5
Task .............................................................................................................6
FUNDAMENTALS .................................................................................................. 7 2.1
2.2
Otto Engine Process ...................................................................................7 2.1.1
Knocking ...............................................................................................8
2.1.2
Thermodynamics ................................................................................11
2.1.3
Compression Ratio .............................................................................17
2.1.4
Thermodynamic Basis for Charging ...................................................19
Charging ....................................................................................................21 2.2.1
2.3 3
4
Turbocharging Nomenclature .............................................................21
Survey of Charging Systems .....................................................................27
BASE ENGINE ..................................................................................................... 38 3.1
Engine Data ...............................................................................................38
3.2
Test Bench Results ...................................................................................39
SIMULATION ....................................................................................................... 42 4.1
Theoretical Background ............................................................................42
4.2
Model Realisation in GT-SUITE ................................................................43
Contents
4.3
5
II
Control-Strategies for 2-Stage Charging Systems ....................................48 4.3.1
Predictive Boost Control .....................................................................55
4.3.2
Boost Control Strategies.....................................................................58
RESULTS ............................................................................................................. 60 5.1
Variation of Boost Control Strategies ........................................................60 5.1.1
Simultaneous Regulated Opening of Compressor- And Turbine
Bypasses; LP Wastegate Also Regulated Before Bypass Opening ...............60 5.1.2
Delayed Opening of Compressor Bypass; LP Wastegate Also
Regulated Before Bypass Opening ................................................................67 5.1.3
Delayed Opening of Compressor Bypass; LP Wastegate Remains
Closed Before Bypass Switching ...................................................................71 5.1.4
Compressor Bypass Remains Closed Throughout Acceleration; LP
Wastegate Remains Closed Before HP Turbine Bypass Opening .................74 5.1.5
Simultaneous Regulated Opening of Compressor- And Turbine
Bypasses; LP Wastegate Remains Closed Before Bypass Switching ...........78 5.2
Stationary Simulation ................................................................................84 5.2.1
5.3
6
7
Load Step 2000 RPM / 2 bar BMEP ...................................................84
Transient Simulations ................................................................................88 5.3.1
Coupled Full-Load Acceleration 25 km/h in 3rd Gear ..........................89
5.3.2
Coupled Full-Load Acceleration 20 km/h in 2nd Gear .........................92
5.3.3
Coupled Full-Load Acceleration 30 km/h in 2nd Gear .........................94
DISCUSSION ....................................................................................................... 98 6.1
Summary ...................................................................................................98
6.2
Outlook ......................................................................................................99
6.3
Conclusions .............................................................................................100
BIBLIOGRAPHY ................................................................................................ 102
List of Figures
III
List of Figures Figure 1.1: Cumulative daily crude oil production from various regions /1/ ....................... 1 Figure 1.2: Trend of reduced CO2 emissions and fuel consumption in major automotive markets /3/ ...................................................................................................... 2 Figure 2.1: Influence of state and relative air-fuel ratio on laminar flame speed /9/ .......... 8 Figure 2.2: Effect of residual gas fraction and relative air-fuel ratio on laminar flame speed (normalised) /9/ ...................................................................................................... 9 Figure 2.3: Cylinder pressure traces showing normal combustion and varying degrees of knocking /10/ ................................................................................................. 10 Figure 2.4: p-V and T-s diagrams of ideal Otto cycle /9/ ................................................. 11 Figure 2.5: Thermal efficiencies of various reference cycles as functions of compression ratio /9/ ....................................................................................................... 12 Figure 2.6: Difference between reference constant volume cycle and real process /9/ .. 12 Figure 2.7: Effect of expansion pressure trace due to dissociation /9/ ............................ 13 Figure 2.8: Variation of specific heat ratio as function of relative air-fuel ratio (referenced to air) /9/....................................................................................................... 14 Figure 2.9: Dependence of specific heat ratio upon pressure and temperature at combustion-relevant conditions /9/.................................................................................. 14 Figure 2.10: Comparison of efficiencies between ideal constant volume reference cycle and fuel-air cycle with dissociation /9/ .................................................................... 15 Figure 2.11: Gas-exchange loop at unthrottled operation /11/ ........................................ 16 Figure 2.12: Gas-exchange loop at low-load, throttled operation /11/ ............................ 16 Figure 2.13: Major components of a modern automotive exhaust gas turbocharger /11/ .................................................................................................................................. 22 Figure 2.14: Determination of A/r ratio in volute housing /14/ ......................................... 22 Figure 2.15: Effect of increasing A/r ratio on turbine mass flow /14/ ............................... 23 Figure 2.16: Compressor map showing operating boundaries such as surge- and
List of Figures
IV
choke lines, as well as lines of constant shaft speed and isentropic efficiency /14/ ....... 24 Figure 2.17:Nomenclature of impeller diameters /15/ ..................................................... 26 Figure 2.18: Effect of impeller wheel trim on a compressor operating map /13/ ............. 26 Figure 2.19: Schematics of engines with single exhaust gas turbocharger (left) and mechanical supercharger (right) ..................................................................................... 28 Figure 2.20: Schematic of a parallel arrangement of 2 turbochargers ............................ 29 Figure 2.21: Schematic of a layout employing 2 turbochargers in series ........................ 30 Figure 2.22: Schematic showing a layout with a combination of a mechanical compressor and exhaust gas turbocharger in a 2-stage arrangement ........................... 31 Figure 2.23: Qualitative map outlines of differently-sized compressors with engine aspirating curves superimposed ..................................................................................... 32 Figure 2.24: Full-load BMEP curves depending on turbocharger selection: red – target; green – large turbocharger; blue – small turbocharger ........................................ 33 Figure 2.25: Transient response of engine load due to varying moments of inertia of turbocharger rotating group /27/...................................................................................... 34 Figure 2.26: Schematic of a parallel sequential or “register” arrangement of 2 turbochargers .................................................................................................................. 35 Figure 2.27: Schematic of a serial sequential or “R2S” arrangement of 2 turbochargers .................................................................................................................. 36 Figure 2.28: Dual-turbocharger arrangement with possibility of switching between serial- and parallel modes (US Patent 7165403 (2007)) /24/ .......................................... 37 Figure 3.1: SGT full-load model calibration – output, BSFC, relative air-fuel ratio /27/ ... 39 Figure 3.2: SGT full-load model calibration – indicated, pumping, friction and brake mean effective pressures /27/ ......................................................................................... 40 Figure 3.3: SGT full-load model calibration – max. cylinder pressure, crank angle of max. pressure, IMEP (high pressure loop), gross indicated efficiency /27/ .................... 40 Figure 3.4: SGT full-load model calibration – air flow, intake manifold pressure, intake manifold temperature, volumetric efficiency /27/ ............................................................. 41 Figure 3.5: SGT full-load model calibration – Intake port and manifold gas dynamics /27/ .................................................................................................................................. 41
List of Figures
V
Figure 4.1: Map of engine model in GT-SUITE ............................................................... 43 Figure 4.2: GT-SUITE model map of turbocharger group subassembly ......................... 44 Figure 4.3: Schematic of two-stage turbocharging (“GT²”) .............................................. 45 Figure 4.4: Model map of coupled engine and vehicle groups /28/ ................................. 46 Figure 4.5: Detailed view of vehicle subassembly model /28/ ......................................... 46 Figure 4.6: PID controller behaviour showing output signal as sum of proportional, integral and derivative parts ............................................................................................ 49 Figure 4.7: PID controller response characteristics depending on chosen gains /29/ .... 49 Figure 4.8: Schematic of control system for 2-stage turbocharging by BMW /24/ .......... 51 Figure 4.9: Controller subassembly for high pressure compressor bypass .................... 52 Figure 4.10: 3-D map of switching mode between 2-stage and single-stage charging ... 53 Figure 4.11: Controller subassembly for low pressure turbine wastegate ...................... 54 Figure 4.12: Behaviour of feedforward controller ............................................................ 56 Figure 4.13: Model map showing the predictive feedforward of the low pressure turbine wastegate controller ............................................................................................ 57 Figure 4.14: Schematic showing the implementation of the boost control system in the vehicle application ..................................................................................................... 58 Figure 5.1: Turbocharger speeds and diameters of wastegate and bypasses; simultaneous bypass actuation; regulated wastegate before bypass opening ............... 61 Figure 5.2: Manifold pressures and BMEP; simultaneous bypass actuation; regulated wastegate before bypass opening .................................................................................. 62 Figure 5.3: Turbine powers; simultaneous bypass actuation; regulated wastegate before bypass opening .................................................................................................... 62 Figure 5.4: PMEP, BMEP, gross IMEP and exhaust manifold pressure; simultaneous bypass actuation; regulated wastegate before bypass opening ..................................... 63 Figure 5.5: Residual gas fraction as function of exhaust manifold pressure; simultaneous bypass actuation; regulated wastegate before bypass opening ............... 64 Figure 5.6: HP compressor operating points; simultaneous bypass actuation; regulated wastegate before bypass opening .................................................................. 65 Figure 5.7: LP compressor operating points; simultaneous bypass actuation;
List of Figures
VI
regulated wastegate before bypass opening .................................................................. 66 Figure 5.8: HP and LP turbine operating points; simultaneous bypass actuation; regulated wastegate before bypass opening .................................................................. 66 Figure 5.9: Turbocharger speeds and diameters of wastegate and bypasses; delayed opening of compressor bypass; regulated wastegate before bypass opening ............... 67 Figure 5.10: Manifold pressures and BMEP; delayed opening of compressor bypass; regulated wastegate before bypass opening .................................................................. 68 Figure 5.11: HP compressor operating points; delayed opening of compressor bypass; regulated wastegate before bypass opening ..................................................... 69 Figure 5.12: LP compressor operating points; delayed opening of compressor bypass; regulated wastegate before bypass opening .................................................................. 69 Figure 5.13: HP and LP turbine operating points; delayed opening of compressor bypass; regulated wastegate before bypass opening ..................................................... 70 Figure 5.14: Turbocharger speeds and diameters of wastegate and bypasses; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening ............................................................................................................... 71 Figure 5.15: Manifold pressures and BMEP; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening ....................................... 72 Figure 5.16: HP compressor operating points; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening .......................... 73 Figure 5.17: LP compressor operating points; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening ....................................... 73 Figure 5.18: HP and LP turbine operating points; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening .......................... 74 Figure 5.19: Turbocharger speeds and diameters of wastegate and bypasses; compressor bypass always closed; wastegate remains closed before HP turbine bypass opening ............................................................................................................... 75 Figure 5.20: Manifold pressures and BMEP; compressor bypass always closed; wastegate remains closed before HP turbine bypass opening ....................................... 76 Figure 5.21: HP compressor operating points; compressor bypass remains closed; wastegate remains closed before HP turbine bypass opening ....................................... 76
List of Figures
VII
Figure 5.22: LP compressor operating points; compressor bypass remains closed; wastegate remains closed before HP turbine bypass opening ....................................... 77 Figure 5.23: HP and LP turbine operating points; compressor bypass remains closed; wastegate remains closed before HP turbine bypass opening ....................................... 77 Figure 5.24: Turbocharger speeds and diameters of wastegate and bypasses; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening ........................................................................................................................... 79 Figure 5.25: Manifold pressures and BMEP; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening ....................................... 80 Figure 5.26: HP compressor operating points; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening ....................................... 80 Figure 5.27: LP compressor operating points; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening ....................................... 81 Figure 5.28: HP and LP turbine operating points; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening ....................................... 81 Figure 5.29: Summary plot of engine speeds of all investigated control strategies ........ 82 Figure 5.30: Summary plot of vehicle speeds of all investigated control strategies ........ 83 Figure 5.31: Summary plot of vehicle acceleration of all investigated control strategies......................................................................................................................... 83 Figure 5.32: Comparison of constant speed load-step performance (red – GT²; blue – . 85 SGT)................................................................................................................................ 85 Figure 5.33: Compressor operating points for SGT; load step from 2 bar BMEP / 2000 RPM to full-load at constant engine speed. ........................................................... 86 Figure 5.34: Turbine operating points for SGT; load step from 2 bar BMEP / 2000 RPM to full-load at constant engine speed. .................................................................... 87 Figure 5.35: HP and LP compressor operating points for GT²; load step from 2 bar BMEP / 2000 RPM to full-load at constant engine speed. .............................................. 87 Figure 5.36: HP and LP turbine operating points for GT²; load step from 2 bar BMEP / 2000 RPM initial condition to full-load at constant engine speed. ................................... 88 Figure 5.37: Engine speeds; acceleration from 25 km/h in 3rd gear ................................ 89 Figure 5.38: Turbocharger shaft speeds; acceleration from 25 km/h in 3rd gear ............. 90
List of Figures
VIII
Figure 5.39: Intake manifold pressure; acceleration from 25 km/h in 3rd gear ................ 91 Figure 5.40: Compressor operating points; acceleration from 25 km/h in 3rd gear ......... 91 Figure 5.41: Turbine operating points; acceleration from 25 km/h in 3rd gear ................. 92 Figure 5.42: Full-load torque and power; 3.5L naturally-aspirated benchmark engine ... 93 Figure 5.43: Comparison of vehicle speed and acceleration, second gear from 20 km/h ................................................................................................................................ 94 Figure 5.44: Comparison of vehicle speed and acceleration, second gear from 30 km/h ................................................................................................................................ 95 Figure 5.45: Comparison of full-load acceleration of FEV GT² from 30 km/h in second gear vs. competitors ........................................................................................................ 95 Figure 5.46: Compressor operating points; acceleration from 30 km/h in 2nd gear ......... 96 Figure 5.47: Turbine operating points; acceleration from 30 km/h in 2nd gear ................ 97
List of Tables
IX
List of Tables Table 3.1: Technical data of FEV SGT (Spray Guided Turbo) engine ............................ 38 Table 4.1: Vehicle-related parameters in GT-SUITE model ............................................ 47
Acronyms and Symbols
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X
Acronyms and Symbols
XI
Acronyms ACEA
Association des Constructeurs Européens d'Automobiles (European Automobile Manufacturers Association)
BDC
Bottom dead centre
BMEP
Brake mean effective pressure
CAC
Charge air cooler
CFD
Computational fluid dynamics
DPF
Diesel particulate filter
EGR
Exhaust gas recirculation
EVC
Exhaust valve closing
EVO
Exhaust valve opening
FMEP
Friction mean effective pressure
GDI
Gasoline direct injection
HP(C)/(T)
High pressure (compressor)/(turbine)
IMEP
Indicated mean effective pressure
IVC
Intake valve closing
IVO
Intake valve opening
LNT
Lean-NOx trap
LP(C)/(T)
Low pressure (compressor)/(turbine)
MPG
Miles per gallon
NA
Naturally-aspirated
NOx
Oxides of nitrogen
OPEC
Organization of the Petroleum Exporting Countries
PI(D)
Proportional-integral(-derivative) (controller)
PMEP
Pumping mean effective pressure
PWM
Pulse-width modulation
Acronyms and Symbols
R2S
“Regulated 2-Stage” turbocharging
RPM
Revolutions per minute
SCR
Selective catalytic reduction
TC
Turbocharger
TDC
Top dead centre
VD
Variable diffuser (compressor)
VTG
Variable turbine geometry
Symbols A
Volute cross-sectional area [m²]
Afront
Frontal area [m²]
ax
Longitudinal acceleration [m/s²]
Cp
Specific heat capacity at constant pressure [J/kgK]
Cv
Specific heat capacity at constant volume [J/kgK]
cw
Drag coefficient [-]
d
Diameter [m]
ei
Inertial mass factor [-]
F
Force [N]
froll
Rolling resistance coefficient [-]
g
Gravitational acceleration [9.80665 m/s²]
HG
Mixture heating value (internal mixture formation) [J/kg]
HU
Fuel lower heating value [J/kg]
i
Factor to denote 2-stroke (1) or 4-stroke (0.5) cycle [-]
I
Moment of inertia [kgm²]
LST
Stoichiometric air-fuel ratio [-]
XII
Acronyms and Symbols
M
Moment [Nm]
) m (m
Mass (mass flow rate) [kg] ([kg/s])
)A mA ( m
Mass (mass flow rate) of exhaust gas [kg] ([kg/s])
)B mB ( m
Mass (mass flow rate) of fuel [kg] ([kg/s])
Md
Brake torque [Nm]
)L mL ( m
Mass (mass flow rate) of air [kg] ([kg/s])
n
Engine speed [1/sec]
p
Pressure
Pe
Brake power [kW]
pme
Brake mean effective pressure (BMEP) [bar]
pmi
Indicated mean effective pressure (IMEP) [bar]
pmr
Friction mean effective pressure [bar]
Pr
Friction power [kW]
pu
Ambient pressure [bar]
q*
Cut-off ratio [-]
q A(,V)
mass-specific heat rejection (at constant volume) [J/kg]
qB(,V)
mass-specific heat addition [J/kg]
QV
Heat addition at constant volume [J]
r
Radius [m³]
rdyn
Dynamic tire radius [m]
RL
Gas constant for air [J/kgK]
s
Entropy [J/K]
t
Time [s]
T
Temperature [K/°C]
v
Velocity [m/s]
XIII
Acronyms and Symbols
V
Volume [m³]
Vc
Clearance volume per cylinder [m³]
Vh
Swept volume per cylinder [m³]
VH
Engine swept volume displacement [m³]
w BL
Laminar flame speed [cm/s]
wC(,s)
Compressor work (isentropic) [kJ]
WT(,s)
Turbine work (isentropic) [kJ]
Angular acceleration [rad/s²]
Compression ratio [-]
Efficiency [-]
C(,s )
Compressor efficiency (isentropic) [-]
e
Brake (also effective) efficiency [-]
i
Indicated (also internal) efficiency [-]
i ,HD
Gross indicated efficiency (high pressure loop) [-]
T (,s )
Turbine efficiency (isentropic) [-]
th(,V )
Theoretical efficiency (constant volume process)
V
Efficiency of constant volume cycle with dissociation [-]
Angle of road inclination [°]
Specific heat ratio (also isentropic exponent coefficient) [-]
Relative air-fuel ratio [-]
a
Volumetric efficiency [-]
C ,HP
Pressure ratio, high pressure compressor [-]
Density [kg/m³]
r
Residual gas fraction [-]
Angular velocity [rad/s]
XIV
Introduction
1
1 Introduction 1.1 Motivation Although at the time of this writing spot prices for crude oil in commodity markets are off their peaks at over USD$147 per barrel set in July 2008, a longer-term and more ominous trend is coming into view. Figure 1.1 shows the cumulative daily crude oil production from various countries/regions compiled from statistics in the years shown. It shows that production has seen a largely constant increase over much of the 20th century ever since large scale exploration and exploitation of petroleum began. This increasing trend has experienced only short-lived reversals due to geopolitical events (e.g. the OPEC oil embargo in 1973), but has generally remained resilient even through economic downturns except for the Great Depression in the 1930s.
Figure 1.1: Cumulative daily crude oil production from various regions /1/
However, beginning in the late-1990s total production has appeared to reach an apex. The period of the last decade, supported by actual up-to-date statistics, has been
Introduction
2
marked by stagnant production growth and in fact slight but prolonged aggregate decline. Although the future prognosis that forecasts accelerating decline is debatable, the trend of the last full decade that reverses that of the previous 80 years should not be ignored, and gives an indication of future prices of crude oil.
Aside from the supply and price of oil, there have been recent regulatory moves within the framework of climate change and the Kyoto Protocol to curb Greenhouse Gas – specifically CO2 – emissions. In Europe, the ACEA established a voluntary limit for CO2 emissions upon its members at 140 g/km by 2008, a target that has been missed. In part resulting from the failure of voluntary compliance, the European Union has mandated a binding limit of 130 g/km for 2012. Failure to meet this on a corporate average basis will result, principally, in fines amounting to 95 € for each vehicle and each g/km over the limit sold by the company. A further 10 g/km up to the 2012 timeframe is to be achieved by non-car-related measures such as the use of biofuels, tyres and by emission reductions in vans. There are further discussions to extend the 2012 binding limits with further CO2 reductions out to 2020, where a limit of 95 g/km is being proposed. /2/
Figure 1.2: Trend of reduced CO2 emissions and fuel consumption in major automotive markets /3/
Introduction
3
However, the regulatory trend of curbing CO2 emissions – and synonymously reducing fuel consumption – is not only occurring in Europe, but rather in all major automobile markets in the world, e.g. North America, Japan and China, as shown in Figure 1.2.
As a result of the multiple pressures of declining crude oil supplies and government regulation among others, automakers are forced to answer to these realities. This has driven research and development into measures to reduce fuel consumption and, by extension, CO2 emissions. A holistic approach must be taken, which would include vehicle-, drivetrain- and engine-side measures. Briefly, vehicle-side measures have principally focused upon mass reduction, aerodynamic improvements (reduction of the product of drag coefficient and frontal area in spite of increasing vehicle sizes) and energy management of electrical consumers throughout the vehicle /4/. Drivetrain measures include the use of low rolling resistance tyres /5/, minimizing brake drag when not applied /6/, improvement of mechanical efficiency through minimisation of friction /7/ and hybridization /8/.
Engine-side technological drivers have also been highly varied and multi-faceted, so only a small number of these are mentioned here for Otto (spark ignition) and Diesel (compression ignition) engines representing a sample of technologies in production. Diesel
Common-rail direct injection
Variable charging (variable turbine geometry, 2-stage turbocharging)
Exhaust aftertreatment (fuel consumption trade-off of LNT and SCR NOx catalysts; DPF backpressure and regeneration strategies)
Otto
Variable valve timing
Direct injection
Charging
Hybridization
Introduction
4
Other common developments to both Diesel and Otto engines can be captioned under the following:
Combustion development
Charge-motion design
Thermal/friction management
Charging
Downsizing
1.1.1
Trend of Downsizing
Common to both Diesel and Otto engines, the concept of downsizing has been a major development trend because of its effectiveness at reducing fuel consumption. Downsizing has been successful and made possible by integrating a combination of multiple measures listed above, along with maturation of the respective technologies and driving down of costs to more mainstream engine programs from high-end lowvolume ones, with the resulting economies of scale.
Downsizing permits a reduction of fuel consumption, since a smaller engine operates at a higher specific load in order to match the same demanded road load at a given driving operating point in order to overcome rolling resistance, inertia, road inclination and aerodynamic drag:
Fdemand Froll Finertia Finclination Fdrag
Equation 1.1
Fdemand froll (mvehicle m payload ) g (ei mvehicle m payload ) a x Froll
Finertia
(mvehicle m payload ) g sin (θ ) 12 ρ cw Afront v 2 Finclinatio n
Equation 1.2
Fdrag
In this way, the downsized Otto engine operates with less degree of throttling and higher intake manifold pressures. Downsizing can be realised through one or both of reduced cylinder dimensions (smaller bore diameter and/or stroke) or fewer number of cylinders. Operating the engine at higher specific load reduces pumping losses in the low-pressure loop of the engine indicator diagram. Friction mean effective pressures tend to increase in proportion to smaller cylinder bore diameter. At first
Introduction
5
glance, this would seem to be a disadvantage for downsizing. However, in absolute terms, the friction power tends be lower, since the increase in friction mean effective pressure pmr is outweighed by the reduction in swept displacement VH :
Pr = i n pmr VH
Equation 1.3
A naturally-aspirated engine is limited in its output. This limit is established by the amount of oxygen that can be drawn from the ambient atmosphere, inducted into the combustion chamber per working cycle and converted to combustion products by burning of the fuel-air mixture. In series passenger-car engines, the brake mean effective pressure (BMEP) is limited to between 12-14 bar at rated torque. In racing engines, this value could exceed 16 bar. The torque output is related to the BMEP via:
Md =
i pme VH 2
Equation 1.4
Power is related to torque via:
Pe = 2 n Md
Equation 1.5
or
Pe = i n pme VH
Equation 1.6
It follows that the engine output can only be increased through some combination of increasing the BMEP or engine speed if the swept displacement should remain constant or not increase. The subject of charging amidst downsizing focuses attention on increasing pme while VH decreases and n remains relatively constant.
1.1.2
Downsizing Challenges
Downsizing with charging presents an effective means for the reduction of fuel consumption compared to a non-downsized, naturally-aspirated variant. However, charging using an exhaust gas turbocharger traditionally incurs a compromise in transient performance due to the mechanical decoupling and mismatch of flow- and operating characteristics between the turbocharger and the engine. Since the turbocharger and engine are not mechanically connected, the driving of the former is achieved only though a fluid connection of the flowing gases. The operating points of
Introduction
6
the turbocharger are therefore not in fixed correspondence at any given time with that of the engine. The transient response of the charger is decoupled from the engine, governed by inertia of the former‟s own rotating assembly and further delayed by finite-time flow phenomena of the two devices. As a result, a turbocharged engine cannot match the transient response of a naturally-aspirated one. This is worse in downsized and highly-boosted engines, as the turbocharger is more responsible for achieving the output over a larger portion of the operating map. The trade-offs in transient response concerning turbocharged engines are discussed in greater detail in Chapter 2.3.
1.2 Task In this report, fundamental backgrounds of the Otto engine process and turbocharging are first briefly introduced. The development of a coupled detailed engine- and vehicle chassis model in GT-SUITE, particularly the control strategy for a 2-stage turbocharged system to be integrated into the model and investigation of its various possibilities, shall be discussed. The results of the control strategy investigation as well as stationary- and transient simulations using models calibrated to measurement test data will then be presented. This work concludes with an outlook for further research and development potential.
Fundamentals
7
2 Fundamentals 2.1 Otto Engine Process The combustion process in reciprocating engines can be broadly divided into two main classifications:
Compression ignition (Diesel)
Spark ignition (Otto)
The Otto process will be briefly summarised here, but the reader is directed to a large number of texts for more detail on both processes, e.g. /9/ and /10/.
In contrast to Diesel engines, the important distinguishing factors of interest in the Otto process relating to this work are primarily three-fold: 1.
Combustion of the trapped fuel-air mixture in the cylinder is normally initiated
through a timed, external ignition energy source, with combustion taking place following this imposed ignition event. 2.
Although there are exceptions, as far as the current work is concerned,
combustion takes place with the mixture in a homogeneous, pre-mixed state with air. 3.
Load control is achieved by regulating the quantity of trapped charge mixture
in the cylinder at a nominally targeted stoichiometric air-fuel ratio. This is accomplished through the use of a throttle in the intake.
Combustion, the progress of the entire cycle and ultimately the engine output will therefore be influenced by a number of factors: The quantity of inducted air and fuel into the cylinders, the preparation of combustible charge, the state and composition of the in-cylinder contents at the initiation of combustion, as well as the ignition timing.
Fundamentals
2.1.1
8
Knocking
The combustion process is fundamentally dependent upon prevailing in-cylinder conditions and is influenced not only by thermodynamic states of temperature and pressure, but also by bulk charge motion, turbulence levels, fuel properties, relative air-fuel ratio and the presence of diluting species such as residual combustion byproducts. Specifically, the state, mixture composition and fuel employed influence the laminar flame speed, which controls the rate of combustion. Figure 2.1 shows a comparison of laminar flame speed of various fuels as a function of the relative airfuel ratio , where:
mL mB Lst
Equation 2.1
It shows that for a wide range of hydrocarbon and alcohol fuels, the laminar flame speed w BL peaks at a value slightly less than 1; for gasoline, this occurs at around 0.85 (upper graph) and drops off from either side of this value. At the limiting cases – at the rich- or lean ignition limits – w BL is theoretically zero /10/.
Figure 2.1: Influence of state and relative air-fuel ratio on laminar flame speed /9/
Fundamentals
9
Figure 2.2: Effect of residual gas fraction and relative air-fuel ratio on laminar flame speed (normalised) /9/
Figure 2.2 shows in investigations using gasoline fuel the dependence of residual gas fraction upon the laminar flame speed. The laminar flame speed decreases substantially with increased dilution but is only relatively weakly influenced by relative air-fuel ratio and pressure.
A phenomenon in the Otto combustion process, known as detonation or “knocking”, can occur. It is characterized by an audible sound from the engine, which is caused by excitations of the engine structure surrounding the combustion chamber resulting from the impact of high pressure shock waves. These shock waves are generated by high rates of pressure rise of colliding flame fronts, one propagating radially outward from the spark plug and one or more fronts caused by spontaneous ignition of the end-gas ahead of the normally propagating one. The effect on in-cylinder pressure can be seen in the comparison of cylinder pressure trace diagrams of a normally operating engine cycle with knocking ones of varying degrees, as shown in Figure 2.3.
Fundamentals
10
Figure 2.3: Cylinder pressure traces showing normal combustion and varying degrees of knocking /10/
In contrast to the smooth progress of cylinder pressure after firing TDC of normal combustion (Figure 2.3, left), the pressure trace with slight knock (middle) shows ripples. The left and middle pressure traces are result from the same operating point at the knock-limited spark timing; the onset of knock occurs in certain cycles as a result of cyclic variation. As the spark timing is further advanced, intense knock results. This is characterised by a highly jagged pressure trace with large fluctuations and peak amplitudes at high frequencies (right). Left uncontrolled and unmitigated, knocking is not only unpleasant to hear, but in serious cases is damaging to engine parts such as pistons.
A unified analytical model of knocking has not yet been developed, but the contributing mechanisms to its causes and control are broadly understood. Knocking is exacerbated by using fuel of too low octane rating, high temperatures and pressures of the prepared mixture at the beginning of ignition (too high compression ratio), early spark timing, high loads, low engine speeds and too low flame speed influenced by aforementioned factors such as residual gas fraction and relative airfuel ratio. A low flame speed allows greater residence time in the unburned zone ahead of the normal flame front from being quickly consumed before forming a spontaneously igniting end-gas.
Fundamentals
2.1.2
11
Thermodynamics
The ideal Otto cycle is made up of the following thermodynamic processes in sequence in a closed, lossless cycle, the p-V and T-s diagrams of which are illustrated in Figure 2.4: 1. Isentropic compression 2. Isochoric heat addition 3. Isentropic expansion 4. Isochoric heat rejection.
Figure 2.4: p-V and T-s diagrams of ideal Otto cycle /9/
The theoretical efficiency of the Otto (constant volume) cycle can be shown to be a function of the geometric compression ratio:
th ,V 1
1
1
Equation 2.2
where
Vc Vh Vc
Equation 2.3
Figure 2.5 shows the efficiencies of various reference cycles plotted against compression ratio. In all cases, the efficiency increase is asymptotic and reaches a point of diminishing returns at high compression ratios.
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12
Figure 2.5: Thermal efficiencies of various reference cycles as functions of compression ratio /9/
Ignition
Figure 2.6: Difference between reference constant volume cycle and real process /9/
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13
The real engine cycle somewhat deviates from each of the above idealised individual processes. Compression and expansion are not exactly isentropic due to wall heat transfer across the control volume boundary. Heat transfer is more pronounced during expansion as the result of higher temperatures. Heat addition and rejection are neither isochoric nor accomplished by boundary heat transfer as in the ideal processes, but rather by ignition before the minimum volume, combustion of fuel and conversion of the working fluid into burned gases in a finite period of time. The real combustion process results in a diminished efficiency than would be achieved through the idealised heat addition. These effects are illustrated in a comparison of ideal- and real processes in a p-V indicator diagram shown in Figure 2.6.
Figure 2.7: Effect of expansion pressure trace due to dissociation /9/
Furthermore, the composition of the in-cylinder gases does not remain constant during the expansion stroke. Rather, the course of combustion and conversion of working fluid from a mixture of air and fuel into burned products at high temperatures result in dissociation and unsteady gas thermo-chemical properties over the course of expansion. Figure 2.7 illustrates a comparison of the pressure traces with and without accounting for equilibrium chemical composition and dissociation during expansion on a p-V diagram.
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14
Figure 2.8: Variation of specific heat ratio as function of relative air-fuel ratio (referenced to air) /9/
Figure 2.9: Dependence of specific heat ratio upon pressure and temperature at combustion-relevant conditions /9/
Fundamentals
15
The actual fuel-air mixture composition and elevated temperatures also result in changing gas properties. Increasingly rich air-fuel mixtures result in reductions in the isentropic exponent coefficient (analogously the ratio of specific heat capacities at constant pressure- and volume, Equation 2.4), because the gas properties of the mixture deviates away from pure air, as illustrated in Figure 2.8. Figure 2.9 shows that declines with increasing temperatures and, starting from approximately 1500K, also decreases at reduced pressures. The temperature and pressure dependence are the result of dissociation.
Cp
Equation 2.4
Cv
Figure 2.10: Comparison of efficiencies between ideal constant volume reference cycle and fuel-air cycle with dissociation /9/
Real combustion, chemical dissociation and mixture composition result in efficiencies that are lower than the idealised reference cycle. This is illustrated in Figure 2.10, which plots reference cycle efficiencies from Equation 2.2 for 1.3 and 1.4 , together with the efficiencies accounting for the real combustion with dissociation at various values of . At a compression ratio of 10 and assuming a stoichiometric airfuel ratio, the combustion cycle efficiency is some 15 percentage points below that of
Fundamentals
the ideal reference cycle (~45% vs. 60%).
Figure 2.11: Gas-exchange loop at unthrottled operation /11/
Figure 2.12: Gas-exchange loop at low-load, throttled operation /11/
16
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17
The idealised cycle also ignores friction and models heat rejection as boundary heat transfer q A . Gas-exchange is also ignored. The latter is an important factor in the analysis of engine performance at full-load and efficiency at part-load. Figure 2.11 shows the gas-exchange loop for an engine at unthrottled operation. In comparison – particularly at low-loads with Otto engines – intake pumping losses due to throttling are a dominant source of reduced efficiency, as shown by the comparatively large negative work loop (shaded area) in Figure 2.12. The working fluid, which has been converted to combustion products, must be expelled out of the system and a fresh charge introduced for the beginning of each subsequent cycle during this gasexchange phase. Gas-exchange also has a decisive impact on the combustion process and the propensity for knocking, in part through the effectiveness of mixture preparation and homogeneity, residual gas content and end-of-compression temperatures.
Summarising, the actual brake efficiency e of a real engine is diminished from the reference cycle value V (accounting for changing working fluid composition and dissociation) by subtracting losses due to non-ideal combustion BV , leakage U , wall heat transfer W , pumping work during gas-exchange LW and friction R ). This is given in Equation 2.5 below:
i ,HD e V BV U W LW R
Equation 2.5
i
2.1.3
Compression Ratio
While in theory a highest possible compression ratio is desirable for maximized thermal efficiency and minimized fuel consumption, in practise, in an Otto engine it is limited by knocking at high loads as a direct result of high end-of-compression temperatures, since in an isentropic case:
T2 T1 ( 1 )
Equation 2.6
Friction losses scale with increasing compression ratio due to higher gas force loadings at the piston rings and bearings as a result of increased cylinder pressures.
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18
Heat transfer losses also increase with higher compression ratios because of increased combustion chamber surface-to-volume ratio as the clearance volume is reduced according to Equation 2.3. Taken as a whole, there is little benefit in brake thermal efficiency – once the aforementioned losses are taken into account – beyond compression ratios in the range of 14–16:1 for naturally-aspirated engines. In practise, compression ratios for non-boosted production engines range from 9–12:1. This may be less than what is technically feasible and optimal for a given engine design, but may be limited by considerations of available fuel quality – specifically as regards to octane rating – in the markets where the vehicle will be sold. Racing engines, being optimized for maximum performance and that have access to higher octane fuels of better consistency, can have compression ratios of about 15:1 /12/.
In boosted engines, the compression ratio must be reduced compared to a naturallyaspirated counterpart in order to avoid knocking on one hand and also reduce the need for high-load fuel enrichment on the other. Fuel enrichment from the stoichiometric value hinders knocking by increasing the laminar flame speed toward the maximum point, e.g. 0.85 for gasoline as shown in Figure 2.1. Enrichment further impedes knocking by reducing the temperature of the fresh charge through the latent enthalpy of vaporisation of the evaporating fuel. Fuel enrichment at high loads is also employed to protect components like pistons, exhaust valves and turbocharger turbines from operating at critical temperatures. This is partially achieved through the aforementioned charge-cooling effect. Also, the additional fuel beyond the stoichiometric air-fuel ratio does not participate in combustion but rather acts as a diluting mass with a specific heat capacity that reduces process temperatures for a given enthalpy of combustion through rearrangement of the FirstLaw relation for ideal gases with constant-volume heat addition:
T
QV m CV
Equation 2.7
There may be further influences to reduce the compression ratio due to design considerations to limit the peak cylinder pressure, or by being forced during calibration to unfavourably late spark timings for knock control. Such retarded timings increase exhaust gas temperatures and delay the 50% burned mass and peak cylinder pressure points to later crank angles. Due to combustion deviating away
Fundamentals
19
from an isochoric process, the indicated efficiency of the cycle is diminished, in spite of a higher compression ratio.
2.1.4
Thermodynamic Basis for Charging
Beginning with a consideration of basic governing thermodynamic equations, it is possible to illustrate the means by which charging can increase the specific power output of the engine. This increase can then be offset by a reduction in engine swept volume in order to achieve a given target performance. In so doing, the efficiency benefits of downsizing are realised by operating the engine predominantly at aforementioned higher specific loads.
The brake efficiency is equal to the quotient of brake power and the rate of fuel energy, the latter being represented by the product of the fuel flow rate and lower heating value:
e
Pe B Hu m
Equation 2.8
By definition, the fuel mass flow rate is related to the air mass flow rate, relative air fuel ratio and the stoichiometric air-fuel ratio LST according to Equation 2.9 below:
B mL m Lst
Equation 2.9
Rearranging the above two equations and solving for the brake power gives:
L Pe e m
Hu Lst
Equation 2.10
Since the lower heating value and stoichiometric air-fuel ratio are fixed for a given fuel, these have no influence on the power output. Likewise is constrained to the stoichiometric value for the chosen combustion process and operation with a threeway catalytic converter. The brake efficiency cannot be arbitrarily increased in correspondence with the demanded increase in power output. Therefore, the brake power can be achieved primarily by increasing the mass flow rate of the charge air:
L Pe ~ m
Equation 2.11
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20
Extending this concept, the BMEP of an engine can be shown to be the product of brake efficiency, volumetric efficiency and specific mixture heating value:
pme e a HG
Equation 2.12
The specific mixture heating value HG for an engine with internal mixture formation (i.e. direct fuel injection; denoted with the bar) is defined in Equation 2.13:
HG
Hu E Lst
Equation 2.13
The ideal gas law gives the relation of charge density at intake valve conditions (subscript „E‟) as functions of pressure and temperature.
E
pE RL TE
Equation 2.14
Equation 2.15 below results from combining Equations 2.12 through 2.14.
pme e a
Hu pE Lst RL TE
Equation 2.15
As before, changes in fuel properties, and brake efficiency are ruled out. The gas constant of the charge working fluid – being essentially air (EGR and residual gases being neglected) and at the relatively low temperatures and pressures at inlet conditions – does not change. The volumetric efficiency is mainly a function of engine design parameters affecting resonance and ram-effect tuning, as well as valve flow coefficients related to piston area. While there is certainly scope to optimize volumetric efficiency, this is relatively limited and cannot be increased by the corresponding order of magnitude of the target increase of BMEP. Charge temperatures also have a lower boundary established by ambient temperature. Therefore, the BMEP is proportional to the charge pressure at intake conditions:
pme ~ pE
Equation 2.16
Summarising, a simplified thermodynamic analysis can show that the brake power output of an engine is a function of the mass flow rate of the charge air, while the BMEP is directly related to the charge pressure. These parameters form the motivations for charging.
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21
2.2 Charging It has been established from the previous section that the means at our disposal to increase the engine torque output (analogous to BMEP) and power are achieved mainly through respective increases in charging pressure and air mass flow rate inducted into the engine, since there is limited scope to change fuel properties or to significantly increase process- and volumetric efficiencies.
From the earliest development of the internal combustion engine, various means have been devised in order to increase the power output though the aforementioned increase in charge pressure and air mass flow rate. The most straightforward is by the use of an auxiliary piston compressor, which has been successfully used in numerous applications over a wide range of engine sizes for vehicles ranging from motorcycles to ships over the past century.
In 1909, Büchi proposed a system to utilise exhaust gas energy acting upon a turbine to directly drive a compressor in a single self-contained unit, instead of employing a mechanical coupling to the engine. The first successful use of turbochargers did not occur, however, until around 1925. It was the advent of the aircraft gas turbine in the 1940s that enabled major strides to be made in turbocharging. In spite of major advancements in the last 30 years, particularly for automotive applications of turbocharging, the principles have remained essentially the same for the past century /13/.
2.2.1
Turbocharging Nomenclature
In this section, some common terms concerning the subject of turbocharging that recur in the present work are defined and explained. Figure 2.13 shows a turbocharger with major components and flow directions labelled.
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22
Figure 2.13: Major components of a modern automotive exhaust gas turbocharger /11/
Figure 2.14: Determination of A/r ratio in volute housing /14/
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23
A/r – describes a geometric characteristic of all compressor and turbine housings. As Figure 2.14 illustrates, it is defined as the cross-sectional area of the volute at a given point divided by the radius from rotational axis centreline to the centroid of that area. At any considered point in the volute, this A/r ratio is constant. Compressor A/r – Compressor performance is largely insensitive to changes in A/r, but generally larger A/r housings are used to optimize the performance for low pressure ratio applications, while smaller A/r ratios are used where
Pressure Ratio [-]
high boost pressures encountered.
Reduced Mass Flow Rate [(kg/s)·K1/2/kPa] Figure 2.15: Effect of increasing A/r ratio on turbine mass flow /14/ Turbine A/r – Turbine performance is greatly affected by changes in the A/r ratio of the housing. Turbine A/r is used to adjust the flow capacity of the turbine, as illustrated in Figure 2.15. Using a smaller A/r ratio will increase the exhaust gas velocity into the turbine wheel, causing the impeller to rotate faster at lower engine speeds, resulting in a quicker boost rise. This will also tend to increase exhaust backpressure and restrict the maximum power at high RPM. Conversely, using a larger A/r turbine housing will reduce exhaust gas velocity and delay boost rise, but the lower backpressure will give better power output at high engine speeds. The turbine A/r should be selected in consideration of the compromise between low- and high RPM performance.
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24
Boost Threshold – is the minimum engine speed at which there is sufficient exhaust gas flow to generate positive manifold pressure, or boost. More specifically, it is the minimum engine speed at which a target boost pressure can be generated.
Turbo lag – is defined as the time delay of boost response after the throttle is opened when operating above the boost threshold engine speed. Turbo lag is determined by many factors. These include turbocharger size relative to engine size, the state of tuning of the engine, the inertia of the rotating group, turbine efficiency, intake flow losses, exhaust backpressure and design of exhaust manifold.
Surge – is the left hand boundary of the compressor map as shown in Figure 2.16. Operation to the left of this line represents a region of flow instability caused by stall or flow separation from the aerodynamic elements of the impeller. This region can be identified by a periodic fluttering or chaffing noise from the compressor. Continued operation within this region can lead to premature turbocharger failure due to heavy thrust loading /15/.
Figure 2.16: Compressor map showing operating boundaries such as surge- and choke lines, as well as lines of constant shaft speed and isentropic efficiency /14/
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25
Surge is most commonly experienced when one of two situations exist. The first and most damaging is surge under load. It can be an indication that the compressor is too large. In this case, the compressor selection should be changed to a smaller one so as to maintain a margin from the surge limit under all anticipated operating conditions.
Surge is also commonly experienced when the throttle is quickly closed after boosting. This occurs because mass flow is drastically reduced as the throttle is closed, but the turbocharger is still rotating and generating boost. This immediately drives the operating point to the far left of the compressor map. Surge will decay once the shaft speed slows enough to reduce the pressure and move the operating point back into the stable region. This situation is commonly addressed by using a diverter valve. Upon closure of the throttle, this valve opens in order to recirculate airflow from the compressor outlet back to the inlet, either electronically from the engine control unit or manually by exceeding a pressure difference threshold in a pneumatic diaphragm actuator. This maintains a sufficient mass flow rate so that the compressor does not operate in the surge region during closed throttle.
Choke – is the right hand boundary of the compressor map of Figure 2.16. The choke line is characterised by curvature of constant speed lines at right-hand side of the map. In the limiting case, these lines of constant speed become almost vertical. The physical explanation of this is that further increase in compressor rotational speed and pressure ratio do not result in proportionally increased mass flow, since flow through the minimum cross-section reaches sonic velocity. One turbocharger manufacturer defines a further criterion for choke as the point on the right-hand side of the maps where the efficiency drops below 58% /15/. The rapid drop in compressor efficiency past this point increases the outlet air temperature and compressor work, since:
C ,s
T2 s T1 T2 T1
Equation 2.17
With the subscript “L” denoting properties for air and states 1 and 2 referring to the compressor inlet and outlet, respectively, the compressor work can be expressed by:
wC
1
C ,s
C p ,L
p T1 2 p1
L 1 L
1
Equation 2.18
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26
Additionally, the turbocharger‟s durability operating near the choke region is compromised by shaft speeds approaching or exceeding the allowable limit and by high thrust loads. If actual or predicted operation approaches the choke limit, a larger compressor is necessary.
Figure 2.17:Nomenclature of impeller diameters /15/
Figure 2.18: Effect of impeller wheel trim on a compressor operating map /13/
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27
Trim – refers to the ratio of the diameters of the impeller outlet (exducer) and inlet (inducer), defined in Figure 2.17. Compressor- and turbine impellers are generally manufactured in relatively few design and size variants within a given manufacturer. Many more variants are possible simply by machining the desired trim. The trim affects the flow characteristics of the device, with larger trims tending to move the entire operating map toward higher mass flows and vice versa, as shown in Figure 2.18. Different turbocharger manufacturers have different conventions for calculating the trim.
2.3 Survey of Charging Systems The simplest layout of a charged engine is a single exhaust gas turbocharger or compressor mechanically connected to the engine. In the latter, shaft power is used to drive the compressor, and the compressor itself may be a positive displacement type like a Roots or rotary vane blower, an internal compression device such a piston compressor of various forms or screw (Lysholm) device /16/. An aerodynamic machine using a centrifugal impeller may also be mechanically driven by the engine, using an appropriate transmission to multiply engine speed to that required for the centrifugal compressor. Another class of charging devices, known as pressure wave superchargers, saw limited application in series production in the 1980s but does not have any modern series passenger car applications and is not considered further /17/.
The above-listed devices are by no means a comprehensive listing. However, modern automotive applications have converged upon using either exhaust-driven centrifugal (radial-flow) compressor/turbines or mechanical compressors of the Roots or
Lysholm
types.
From
this
point
onwards,
any
exhaust
gas-driven
turbine/compressor pair will be referred to as a turbocharger, while a supercharger will refer to a mechanically-driven device.
In all cases of a charged engine with a single compressor device, ambient air enters into the compressor and is ejected into the intake tract of the engine at higher pressure. Figure 2.19 shows schematics of engines employing a single exhaust gas turbocharger (left) and mechanical compressor (right).
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28
( ) Figure 2.19: Schematics of engines with single exhaust gas turbocharger (left) and mechanical supercharger (right)
Turbochargers themselves can be generally of a wastegated type as shown in schematic form in Figure 2.19 or with variable turbine geometry (VTG, shown in brackets) as means of boost pressure control. IHI has also shown a concept of controlling the flow output of a turbocharger via variable compressor diffuser geometry (VD) technology /18/.
Although not a form of charging in the strict sense of the term, the concept of compounding is briefly explained. Here, a turbine placed in the exhaust of a combustion engine in addition to a turbocharger is mechanically connected to the crankshaft and harnesses exhaust gas energy, which is converted to mechanical work that directly augments the engine output. The ideal, maximum amount of mechanical power in a turbine is equal to the product of the mass flow rate of exhaust gas and the isentropic expansion work across the turbine between the inlet and outlet states. The maximum extractable turbine work is given by Equation 2.19, with subscript “A” denoting properties for exhaust gas, and states 5 and 6 referring to turbine inlet and outlet, respectively:
wT T ,s
p C p , A T5 1 6 p 5
A 1 A
Equation 2.19
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29
The next logical step in charging systems is to employ more than one compressor device. Any combination of turbocharger or supercharger could be employed. In the simplest case, two devices can be laid-out either in series or in parallel and both are always operating simultaneously.
Figure 2.20: Schematic of a parallel arrangement of 2 turbochargers
In the case of a parallel arrangement, the pressure ratios of both devices are the same but the intake and exhaust mass flows are divided equally between each device, as shown in Figure 2.20. The fresh air flow discharged from the two identical compressors at the same operating points maybe be merged together before entering the engine intake manifold or may remain separate to serve one-half the number of cylinder or separate banks of a V- or boxer engine layout, for example.
In a series layout, ambient air enters one of the compressors and the outlet from the first compressor is connected to the inlet of the second. Similarly, the exhaust gases enter the first turbine, and the outlet is connected directly to the inlet of the second. Figure 2.21 shows a schematic of two turbochargers connected in series.
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30
Figure 2.21: Schematic of a layout employing 2 turbochargers in series
The intake- and exhaust mass flows, obeying the continuity equation, are the same between both devices, but the pressure ratios of each stage are multiplied. Summarising, when the pressure ratios are analogously expressed as resistances and mass flow rate as current, the behaviour of series and parallel flow arrangements correspond to an electric circuit analogy using Ohm‟s Law.
Figure 2.22 shows a layout combining a turbocharger and supercharger in a serial arrangement. In 2-stage mode, air is first compressed by the supercharger and is then routed into the inlet of the turbocharger compressor. The supercharger can be disengaged via an electromagnetic clutch; when this occurs, a bypass valve opens in order to allow air to flow around the supercharger compressor.
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31
Figure 2.22: Schematic showing a layout with a combination of a mechanical compressor and exhaust gas turbocharger in a 2-stage arrangement
A qualitative evaluation of compressor maps will illustrate the limits of turbocharger dimensioning and selection to engine performance. Dependencies of air mass flow rate on power output and charge pressure on the BMEP have been derived in Equation 2.11 and 2.16, respectively. An engine operating with representative process- and volumetric efficiencies will have an aspirating curve at full-load, sketched qualitatively in Figure 2.23, that corresponds to the target full-load BMEP (red curve) in Figure 2.24. Due to the differences between the aspirating characteristics
of
an
internal
combustion
engine
and
aerodynamic
compressor/turbine, a small single turbocharger selected to give good transient response and full-load performance at low engine speeds will approach the upper boundary of the compressor map limited by maximum permissible shaft RPM, thus limiting the attainable boost pressure ratio. Also, a small compressor will be bordered on the right hand side of the map by the choke limit, therefore hindering performance at higher engine speeds, as shown in Figure 2.24 (blue curve). The engine output would therefore have to be de-rated to a lower peak BMEP and power output in order to remain within the operating boundary of this chosen turbocharger.
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32
4.0
3.5
Pressure Ratio [-]
3.0
2.5
2.0
1.5
1.0 0.00
0.05
0.10
0.15
0.20
0.25
0.30
Corrected Mass Flow Rate [kg/s]
Figure 2.23: Qualitative map outlines of differently-sized compressors with engine aspirating curves superimposed
Conversely, a turbocharger optimized for a high rated power at high engine speeds must be dimensioned larger. By appropriate matching, a turbocharger can support the necessary mass flow rates in order to reach the target power output levels, while also having sufficient margin to operate at the required pressure ratios in order to reach the BMEP target. However, the large compressor will have the surge line shifted to higher mass flows compared to the small counterpart. Therefore the aspirating curve for the target full-load BMEP at low engine speeds will be in the surge region, where operation must be avoided. The output of the engine will have to be de-rated at low speeds in order that the aspirating curve will clear the surge line with sufficient margin as shown in Figure 2.23 (green curve). The larger turbocharger also has an inherently higher boost threshold RPM, thereby limiting the boost pressure at low engine speeds.
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33
Figure 2.24: Full-load BMEP curves depending on turbocharger selection: red – target; green – large turbocharger; blue – small turbocharger
At the same time, the angular acceleration or time rate of change of angular velocity of the rotating components – compressor- and turbine impellers and shaft – of the turbocharger is given by:
d M dt t I
Equation 2.20
Rearranging Equation 2.20, it can then be shown that the elapsed time required for the angular velocity to change between beginning and ending values is directly proportional to the mass moment of inertia I :
t I
M
Equation 2.21
A larger turbocharger will exhibit worse transient response than a smaller one due to this increase in the moment of inertia, which scales with the square of diameter:
I ~ d2
Equation 2.22
Figure 2.25 shows the transient effects of varying the turbocharger moment of inertia. Decreasing I results in quicker dynamic torque generation of the engine.
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34
Figure 2.25: Transient response of engine load due to varying moments of inertia of turbocharger rotating group /27/
By some rational arrangement of combining a small turbocharger and a large one, the objective is that the benefits of both can be realised in the same engine. Therefore, a small turbocharger allows quick transient response and good performance at the lower end of the engine speed regime. The large turbocharger supports the mass-flow rate and boost pressure ratio for simultaneously high power output and BMEP at higher speeds. Finally, both turbochargers operating simultaneously bridge the operating gaps between both individual units, as shown in the shaded part in Figure 2.24.
Besides full-time parallel- and series dual turbocharger arrangements, sequential systems have been introduced in production since the 1980s. The concept is experiencing renewed interest /19/, /20/. A distinction is made from a parallel sequential system (also referred to in literature as “register” turbocharging that has been previously implemented in gasoline engines) and a relatively recent development of series sequential charging /21/, /22/ known by the concept‟s developers as “Regulated 2-Stage (R2S)” turbocharging. Figure 2.25 shows a schematic of a register turbocharging system, while Figure 2.26 shows R2S counterpart.
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35
Figure 2.26: Schematic of a parallel sequential or “register” arrangement of 2 turbochargers
In this work, the layout of the turbochargers is an adaptation of the series 2-stage charging concept, which is already in volume production for Diesel passenger car applications. The advantages of serial charging include the ability for high boost pressures. This is accomplished without operating each turbocharger stage at particularly high pressure ratios, due to the multiplication effect of pressure when the flow is arranged in series. However, the adaptation of series 2-stage charging for Otto engines has some challenges. In particular, due to load being determined by quantity control of the charge air at a tightly regulated air-fuel ratio, any fluctuation of boost pressure and air mass flow rate will result in a corresponding change in engine output, which could then be felt by the vehicle occupants as an unsmooth dynamic behaviour. This is not so critical in Diesel engines due to quality control of output; fluctuations of boost pressure and mass flow will have less of a detectible effect.
Fundamentals
36
Figure 2.27: Schematic of a serial sequential or “R2S” arrangement of 2 turbochargers
The aforementioned layouts for arranging one or more compressor devices are by no means an exhaustive listing. With two charging devices and various layouts and control strategies, a very large number of potential combinations are possible. Indeed, the degrees of freedom are further multiplied when even more charging devices and bypasses are employed, such as a triple-turbocharged passenger car Diesel concept introduced at the Geneva Motor Show in 2005 /23/. Figure 2.27 shows the schematic from US Patent 7165403 (2007) that employs two turbochargers. By control of as many as five bypass valves, the system can be switched between parallel- and series turbine arrangements along with switchable high- or low-pressure EGR /24/.
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37
Figure 2.28: Dual-turbocharger arrangement with possibility of switching between serial- and parallel modes (US Patent 7165403 (2007)) /24/
Non-mainstream designs include the Hyperbar /13/ and variations of specialised variations of turbocompounding such as differential compounding /25/. So far these have been more of research interest and have not seen widespread series production.
Base Engine
38
3 Base Engine 3.1 Engine Data The base engine used for the test bench investigations and also in the build-up of the demonstrator vehicle is a liquid-cooled, 4-stroke inline 4-cylinder engine developed by FEV Motorentechnik. It is a versatile and modular platform to demonstrate current and emerging engine technologies. Therefore, the base specification, for which 2stage turbocharging is a further development, incorporates 4-valves per cylinder with a pent-roof combustion chamber shape and features spray-guided gasoline direct injection (GDI) with centrally-positioned fuel injectors as well as intake- and exhaust camshaft phasing. The engine is also already laid-out and dimensioned for peak cylinder pressure of 140 bar /26/.
Table 3.1: Technical data of FEV SGT (Spray Guided Turbo) engine Engine Specifications Engine Type Bore Stroke Displacement Compression ratio Cyl. head layout
Turbochargers used in GT² HP Compressor LP Compressor
Inline-4 cylinder 81 mm 87 mm 1.8 L 9.8 (reduced to 8.5 in GT² study) dual overhead camshafts, 4 valves per cylinder
1574 CBK 2283 DBCHA
Intake cam phaser Exhaust cam phaser Peak firing pressure Fuel system pressure Fuel pump Injector type
HP Compressor LP Compressor
60° CA full adj. range 60° CA full adj. range 140 bar max. 200 bar max. 3-piston axial high pressure pump Bosch HDEV4 Piezo, outward opening, central position
KP35-240.82 K04-6.88
Base Engine
39
The important specifications of the FEV SGT (Spray-Guided Turbo) engine, on which the 2-stage concept is based, is summarised in Table 3.1. The 2-stage turbocharging concept has been given the designation “GT²” (Gasoline Turbo, 2-Stage). Therefore, from here GT² will refer to the 2-stage concept and SGT to the original single-stage basis.
3.2 Test Bench Results A detailed model of the SGT engine has been developed in GT-POWER and extensively validated with test bench measurements by Höpke /27/. This validated model forms the basis for the GT² in this current work. Here, the calibration of important base engine parameters is presented. These parameters include full-load brake output, BSFC, friction, pumping work, combustion process and intake gas dynamics.
Figure 3.1: SGT full-load model calibration – output, BSFC, relative air-fuel ratio /27/
Base Engine
40
Figure 3.2: SGT full-load model calibration – indicated, pumping, friction and brake mean effective pressures /27/
Figure 3.3: SGT full-load model calibration – max. cylinder pressure, crank angle of max. pressure, IMEP (high pressure loop), gross indicated efficiency /27/
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41
Figure 3.4: SGT full-load model calibration – air flow, intake manifold pressure, intake manifold temperature, volumetric efficiency /27/
Figure 3.5: SGT full-load model calibration – Intake port and manifold gas dynamics /27/
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4 Simulation 4.1 Theoretical Background Simulation encompasses a very wide field of engineering and science that attempts to build model representations of physical systems employing a combination of fundamental mathematical and physical laws as well as empirical relationships. The model should accurately describe the behaviour of the physical system being modelled. Trends in the simulation should follow trends in the physical system and vice versa. The models can then be computationally evaluated. Once the model is calibrated and validated, simulations can be run to predict outcomes and cases beyond the validated conditions.
The code solves fundamental relations such as Newton‟s laws, conservation of mass, momentum and energy, ideal gas equations and the laws of thermodynamics depending upon the type of simulation being performed (e.g. mechanics, flow, etc.).
Rather than evaluating a whole system in its entirety or as in infinitesimal continuum, these models are usually evaluated in discretized form by means of finite elements or volumes. A finer degree of discretization – or a greater number of elements – can improve calculation accuracy but at the expense of increased computation time. Therefore, a compromise must always be struck between these conflicts.
A discussion of simulation also invariably involves the subject of the dimension or order. This refers to the order on the Cartesian coordinate system. A 3-D simulation evaluates the governing equations in all three Cartesian directions and can therefore describe the states in 3-dimensional space, i.e. a volume. A 3-D CFD simulation is one such example. At the other extreme, a 0-D simulation is spatially invariant. Cylinder pressure trace analysis is a 0-D example. The simulations performed in GTSUITE in this study draw from a combination of 0- and 1-D calculations. Pipe flows in GT-SUITE are evaluated in 1-D (along the flow direction) and map lookups are usually zero-dimensional.
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4.2 Model Realisation in GT-SUITE The model developed for the present work is derived from the coupling of previously separate vehicle model in GT-DRIVE by this author /28/ with a detailed engine model in GT-POWER by Höpke /27/.
Figure 4.1: Map of engine model in GT-SUITE
Figure 4.1 shows a map of the detailed engine model in GT-SUITE. Major engine components and subsystems can be identified. These include the engine itself with
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its cylinders, intake- and exhaust manifolds, turbocharger assembly, charge air cooler and connecting pipes in the intake and exhaust.
HPC-Bypass
HPT-Bypass HP-Turbocharger
LP-Turbocharger
LPTWastegate
Figure 4.2: GT-SUITE model map of turbocharger group subassembly
A view into the turbocharger subassembly, shown in Figure 4.2, also depicts a resemblance of the GT-SUITE model with the schematic in Figure 4.3, with the two turbochargers and associated bypasses. A convention is established in this work to distinguish the terms “bypass” and “wastegate”, as both terms are technically equivalent in meaning. As Figure 4.2 shows, the term “wastegate” shall be associated explicitly with the low pressure turbine, and for the high pressure stage, “bypass” shall be employed – whether for the compressor or turbine. If not otherwise specified, wastegate mentioned alone will refer to the diverting of exhaust gas around the low-pressure turbine; while bypass will correspond to exclusively the high pressure stage and will be distinguished further between the compressor and turbine.
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HPC-Bypass
HPT-Bypass LPT-Wastegate
Figure 4.3: Schematic of two-stage turbocharging (“GT²”)
The previously described engine model is integrated into a vehicle model that represents the chassis into which engine is integrated in reality. Figure 4.4 shows this coupled model with three main interconnecting functional modules, namely Engine, Vehicle and Driver. Additional functionality and interconnections are added over these primary modules as necessary.
The Driver object controls the actuation of the accelerator and brake pedals, respectively, by means of internal PID controllers with closed-loop feedback. In this way, the Driver attempts to follow a desired driving schedule from a transient speedtime profile array by manipulating the respective accelerator and brake inputs. The driver object also determines the speeds at which up- and downshifts occur through a user-defined array as well as the necessary clutch actuation.
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Driver
Engine
Vehicle
Figure 4.4: Model map of coupled engine and vehicle groups /28/
Axle with brakes
Transmission with final drive
Axle with brakes
Axle with brakes
Vehicle body
Axle with brakes
Figure 4.5: Detailed view of vehicle subassembly model /28/
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A closer look in the vehicle subassembly as shown in Figure 4.5 reveals the vehicle body and drivetrain comprising the transmission, driveshaft, final drive, axles and brakes. The model employs rigid connections between functional components and all rotating objects have non-zero moments of inertia.
The reader is directed to refer to /28/ for a more detailed explanation of theoretical background and determination of the chassis- and vehicle-side parameters like aerodynamic and coast down coefficients.
Table 4.1: Vehicle-related parameters in GT-SUITE model Description Ambient pressure Ambient temperature Drag coefficient 1st gear ratio 2nd gear ratio 3rd gear ratio 4th gear ratio 5th gear ratio 6th gear ratio Final drive ratio froll - constant part froll - linear part Tire dynamic radius
Unit Value Description bar 1 Vehicle curb mass °C 20 Payload mass 0.3494711 Frontal area 3.385 1st gear moment of inertia 2.05 2nd gear moment of inertia 1.433 3rd gear moment of inertia 1.088 4th gear moment of inertia 0.868 5th gear moment of inertia 0.7 6th gear moment of inertia 4 Final drive moment of inertia 0.0100136 Driveshaft inertia h/km -7.69E-07 Axle inertia (per corner) mm 312 Engine inertia
Unit kg kg m² kg·m² kg·m² kg·m² kg·m² kg·m² kg·m² kg·m² kg·m² kg·m² kg·m²
Value 1350 100 2 0.015 0.015 0.015 0.015 0.015 0.015 0.015 0.015 0.8 0.15
The important vehicle parameters to be inputted into the GT-SUITE model are summarised in Table 4.1.
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4.3 Control-Strategies for 2-Stage Charging Systems The traditional method of modelling the load regulation and boost pressure control system in a steady-state engine model in GT-POWER is usually adequately served by a simple PID controller (actually PI in practice, since the differential part is not used).
The following equations are solved in the PID controller:
dx1 u dt
Equation 4.1
dx 2 u x2 dt
Equation 4.2
K K y K p D u K I x1 D x2
Equation 4.3
where K P is the Proportional Gain, K I is the Integral Gain, K D is the Derivative Gain, is the Derivative Time Constant, y is the controller output, u is the difference between the Reference Signal value and the input signal value, and x1 and x 2 are the state variables /29/.
The controller has constant gain parameters for a given engine load and speed within defined actuation limit values acting upon the actuator. This system acts as a closed control loop with feedback, whereby the controller output represents a transfer function of an input disturbance, which is the instantaneous difference of the sensed input to the target value.
However, in a transient model with engine speed changing, a control-loop system like that described above will not be adequate because of the highly complex, unsteady and dynamic nature of the system to be controlled, itself part of other interconnecting and interacting systems. Consider for example a control loop for boost pressure of a single turbocharger by actuating a wastegate.
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Figure 4.6: PID controller behaviour showing output signal as sum of proportional, integral and derivative parts
Figure 4.7: PID controller response characteristics depending on chosen gains /29/
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As long as the disturbance signal is a negative result, i.e. the actual sensed boost pressure is below the target, the proportional output of the PI controller will be negative. The controller output signal begins to open the wastegate only when the sum of the proportional, integral and derivative output exceeds the defined lower limit, as shown in Figure 4.6. In the case of a wastegate, it can only operate between full closure and full opening areas value. Due to the inertia of the turbocharger rotating group as well as gas flow dynamics, there will tend to be a temporal lag between the wastegate actuation and the actual response in boost pressure / engine torque output.
This lag tendency of the wastegate actuation results in an overshoot in boost pressure (controlled variable). This is followed by overcompensation by the controller, and then undershoots as a consequence. The characteristic of a control system using a PI controller is for the controlled variable to therefore oscillate around the target before converging to it (Figure 4.7 (a)). In the worst cases, the magnitudes of over-/undershoots do not converge (b) or rather even diverge, or the time at which the controlled variable reaches the target value is unacceptably long ((c) and (e)). These characteristics can be optimized by selecting appropriate controller gain parameters and matching these parameters with the inherent delays in the sensed input and response of the physical actuator.
In a control loop with no lag, a properly optimized controller can reach a stable convergence at the target final value relatively quickly and with minimal over- or undershoots. However, modelling transient boost control in an engine with changing speed introduces significant lag between wastegate actuation and boost pressure in response. Even the best optimization of a simple PI-based closed-loop feedback control system will result in a compromise between stability/damping of the fluctuations and the convergence time. At a minimum, the integral gain K I is timedependent; therefore, the time constant is set as a function of the engine speed.
Furthermore, the control system of a modern combustion engine must balance the simultaneously interacting effects of numerous controlled- and uncontrolled systems of physical phenomena. For example, the boost pressure response is influenced not only by wastegate position but also by turbocharger inertia, fluid dynamics of the intake/exhaust flow loops and, especially at part-load, throttle position and exhaust
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gas recirculation (EGR). More decisively, brake engine torque will not only be a function of boost pressure, but also the effects of exhaust backpressure upon pumping work during gas exchange and residual gas fraction upon the combustion process. These highly complex interactions must be brought under control amidst a highly dynamic system. In the case of the 2-stage system in this study, the boost control is further complicated by the introduction of additional degrees of freedom in the form of the second turbocharger and of bypass valves around the HP compressor and turbine, which must be regulated with separate control systems. The simultaneous coordination of multiple control systems must be done in such a way that their effects minimise the interference upon one another. Anti-Reset-Windup Parameter Adjustm. Actual Boost HP Turbine Bypass Signal Target Boost PI Controller Limiter
Boost Pressure Sensor Pre-control Map at LP Turbocharger
Pre-control Map for HP Turbocharger
Controller for the HP Turbine Bypass
Chargecooler HPC Bypass Signal
Torque
HPC Bypass
Engine Speed
Gear Base Map HPC Bypass Hysteresis v
Switching Logic
Air Pressure Corr. Map
HP Turbocharger
HP Turbine Bypass Flap
Controller for the HP Compressor Bypass
Air Pressure Air Pressure Correction
LP Turbocharger Anti-Reset-Windup Parameter Adjustm. Actual Boost
LP Turbine Wastegate
LP Turbine Wastegate Signal
AIR
EXHAUST GAS
Target Boost PI Controller Limiter
Pre-control Map at LP Turbocharger
Pre-control Map for HP Turbocharger
Controller for the LP Wastegate
Figure 4.8: Schematic of control system for 2-stage turbocharging by BMW /24/
The general layout of the boost control model in this work approximates the approach taken from the series production 2-stage turbocharged Diesel engines by BMW as shown in Figure 4.8 /24/. In the upper window for the control of the HP turbine bypass flap, the sensed boost pressure is compared against a stored map of requested boost pressure as functions of engine load and speed.
The difference of the two
values are then inputted into a PI controller, whose parameters such as limits and gains as previously described are also mapped based on engine load and speed.
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What is different from the aforementioned simple feedback loop is the addition of two feedforward loops. One adds a mapped augmenting or pre-conditioning signal to the controller output, while the other switches between the controller and an imposed mapped value for the actuator position based on a trigger signal from a separate switching logic in the second window. The switching logic is a function of variables such as the state of HP compressor bypass, vehicle speed and selected gear. The feedforward-imposed value acts as an initial condition for the system, which results in quicker matching of the target and actual boost pressure. This is because the controller must only operate through a narrower range of actuation. The second switched feedforward loop is activated when the compressor bypass is open. When this happens, the turbochargers no longer operate in 2-stage mode. Rather, the HP turbocharger essentially idles and the boost pressure and mass flow are both driven by the LP turbocharger. The control system of the HP turbine bypass and LP turbine wastegate are schematically very similar as can be seen in the top and bottom windows of Figure 4.8.
Min. limit bef ore mode switch
Max. limit bef ore mode switch
Bypass map PI controller
Min. limit af ter mode switch
Max. limit af ter mode switch
Figure 4.9: Controller subassembly for high pressure compressor bypass
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The boost control strategy developed for this thesis follows a broadly similar approach to that described for the BMW system shown in Figure 4.8. Figure 4.9 shows a GT-SUITE model map for the HP compressor bypass control system. At the heart of this control system is a PI controller, which evaluates the difference of intake manifold pressure to a map of target boost pressure as a function of engine load and speed.
The controller also receives upper and lower limits of actuation from a map that is a function of load and speed, as shown in Figure 4.10. The map gives a binary output – 0 for bypass closed and 1 for open – that triggers switches that activate different signal constants for the bypass orifice diameter based on the operating mode.
Figure 4.10: 3-D map of switching mode between 2-stage and single-stage charging
The HP turbine bypass employs 2 separate maps for opening and closing, the difference between both maps being an offset of the engine torque. This gives a hysteresis effect in order to prevent oscillation of the bypass near the switching operating point boundary. The controller output signal is conditioned with a first order filter with a chosen time constant to reflect the physical wastegate actuator‟s finitetime reaction ability.
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Distinct from the BMW system, where the HP turbine bypass is regulated with a PI controller while the compressor bypass is switched, in the present work the control systems of HP turbine- and compressor bypasses are both regulated and architecturally identical. Switched actuation of the HP compressor bypass is simply a special case of the regulated control system, so that either switching or regulation can be implemented without changing the architecture of the control system in the GT-SUITE model. This results in an additional degree of freedom that allows for flexibility in investigations of control strategies of both turbochargers, as will be discussed in a later chapter.
Min. limit bef ore mode switch
Max. limit bef ore mode switch
From switching logic
PI controller
Min. limit af ter mode switch
Max. limit af ter mode switch Predictive f eedf orward
Figure 4.11: Controller subassembly for low pressure turbine wastegate
Also different from the BMW system is that the control of the wastegate is with a PI controller evaluating engine load as the input signal against a reference load instead of boost pressure. This was chosen because if multiple PI controllers are working with the same boundary parameters and input signal, i.e. intake manifold pressure, the outputs will also be the same. Contentions can therefore arise in a system with multiple controllers. Furthermore, no prior mapping of wastegate positions, boost
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pressure, etc., of the engine had been conducted on a test bench before the build-up of the GT-SUITE model. As a result, measurement data was not available to build a target boost pressure map calibrated to the correspondence of boost pressure and output of the actual engine.
4.3.1
Predictive Boost Control
The work of this thesis also attempts to address the limitation of the PI controllerbased system‟s ability to deal with lagged response systems as described in the previous chapter. A predictive, pre-emptive feedforward to the basic control architecture from the previous section has been conceived for the actuation of the LP wastegate. This has been found to reduce over-/undershoots in boost pressure and result in a more constant engine output during simulated full-load acceleration runs. Its motivation and function is described in the following. Figure 4.11 shows the GTSUITE model map of the wastegate control system. Similar to the BMW system, the function of the wastegate is activated by a switch triggered by the state of the bypass valve. The PI controller receives the engine load directly as its sensed input. This is compared against a reference load, which is a mapped value based on accelerator pedal position – normalised to the engine full-load torque curve – and engine speed.
The predictive feedforward forecasts boost – more specifically brake torque – overshoots within a defined window of the target; in this case the feedforward is active +/- 2 bar BMEP of the instantaneous target value. This is in order to minimise the interference of the feedforward on the rest of the control loop and the PI controller. The logic behind the predictive controller is to determine the trajectory of the input signal (BMEP) before it reaches the target. If the input signal is below the target but is approaching it at a high trajectory or rate, it is a reasonable prediction that an overshoot will occur, even before the threshold is reached. With this prediction the feedforward can pre-emptively take the necessary actions to damp out the trajectory of the input supplemental to the main controller, either by adding an augmenting signal to open the wastegate further than the PI controller signal itself, or also by subtracting the signal of the controller, subject to the actuation limits of the wastegate diameter.
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Figure 4.12: Behaviour of feedforward controller
The behaviour of the feedforward can be seen in Figure 4.12. The output from the PI controller is shown in the blue curve. The feedforward signal (green curve) adds or subtracts from the output of the PI controller. The actual actuated signal (red) is the sum of the feedforward and PI controller outputs, within the defined minimum and maximum limits. During conditions that require a spike in the wastegate area such as at ~t=1.8 sec., for example during a boost pressure overshoot, it can be clearly seen that the feedforward opens the wastegate earlier than the PI controller would itself. As the input error declines and converges to the target value, the feedforward signal approaches zero and the PI controller regulates the system. The feedforward inherently disturbs the intended signal of the PI controller; therefore the feedforward gain should remain at a modest value relative to the optimised proportional and integral gains. Because of the interaction of the feedforward with the PI controller, the parameters of the latter should be re-optimised. In investigations in this paper, the PI controller gains were left constant with or without the feedforward enabled.
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Figure 4.13: Model map showing the predictive feedforward of the low pressure turbine wastegate controller
The implementation of the boost controller in the actual vehicle, as illustrated schematically in Figure 4.13, is similar to the BMW approach and that implemented in the GT-SUITE model, to the extent that the HP turbine bypass valve and wastegate are regulated with PID controllers comparing boost pressure to mapped requested values or set points as functions of engine speed and load. There are also switches that toggle between regulated mode via the PID controllers or feedforward from the mapped pre-control of boost pressure set points. These are triggered by a controller prioritization logic, which controls the compressor bypass state (open or closed). This prioritization permits only one of either the HP turbine bypass or LP wastegate to operate at any given time. The investigations of boost control strategies in the next section will consider both this approach as well as allowing both the wastgate and respective bypasses to operate simultaneously.
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Pre-control of Boost Pressure Set Point
Controller Prioritization
Actual Boost Pressure PID-controller HP turbine
PID-controller LP turbine
Actual Boost Pressure
True = Compressor Bypass Open
Duty Cycle: 0% = open Duty Cycle: 100% = closed
Figure 4.14: Schematic showing the implementation of the boost control system in the vehicle application
4.3.2
Boost Control Strategies
An objective of this thesis is to investigate different control strategies of the 2-stage turbocharging system. In particular, it relates to the different ways to regulate the bypasses and wastegate and the resulting behaviour of the individual turbocharger operating points and overall engine performance. Five strategies of switching from serial 2-stage mode to LP stage-only mode in the context of transient full-load acceleration of the coupled engine and vehicle models in GT-SUITE are investigated: Simultaneous regulated opening of compressor- and turbine bypasses; LP wastegate also regulated before bypass opening; Delayed opening of compressor bypass; LP wastegate also regulated before bypass opening; Delayed opening of compressor bypass; LP wastegate remains closed before bypass switching; Compressor bypass remains closed for the whole acceleration process; LP wastegate remains closed before bypass switching; Simultaneous regulated opening of compressor- and turbine bypasses; LP wastegate remains closed before bypass switching.
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Each of the above strategies results in different turbocharger operating points during simulations involving full-load acceleration and also different degrees of smoothness of the dynamic torque generation. The operating points and progress of turbocharger speeds in a transient context are of interest in order to ensure that that turbocharger will not operate at a critical condition (for example overspeed) or in sustained compressor surge. The results of these investigations will be elaborated upon in Section 5.1.
This section has described the control architecture of the 2-stage turbocharging system, from a description of a system in current series production from which the present work is inspired, to the comparison of similarities and differences with the model in GT-SUITE employed in this work, and the realisation in a vehicle application. This section has also introduced various strategies of controlling the turbocharger bypasses with the goal of ensuring the smoothest possible engine torque output and understanding their influence upon the transient characteristics and critical operating points.
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5 Results In this chapter, the results of simulation investigations performed in GT-SUITE shall be described. The first section presents the effects of varying the boost control strategy by actuation of the bypasses and wastegate upon the operating points of both HP- and LP compressors and turbines. The second section covers a stationary load step from 2000 RPM / 2 bar BMEP. This is compared against the single-stage turbocharged SGT, from which the two-stage GT² development is derived. This assists in comparing the base engine model of the GT² to the already validated SGT. This is followed in the third section by transient simulation results of the coupled engine- and vehicle model in GT-SUITE. Various starting vehicle speeds and selected gears are simulated. A comparison of the GT² concept to various other benchmarked examples is made using a combination of simulation and in-car measured results. Operating point traces of the compressor- and turbine maps are plotted for the investigated cases.
5.1 Variation of Boost Control Strategies As introduced in Section 4.3, the GT-SUITE model was built-up to allow investigation of different boost control strategies and to study their effects in a transient context on turbocharger operating points and dynamic torque generation. The results of five investigated cases are presented in turn. In all cases in this section, full-load acceleration from 20 km/h in second gear is investigated.
5.1.1
Simultaneous
Regulated
Opening
of
Compressor-
And
Turbine
Bypasses; LP Wastegate Also Regulated Before Bypass Opening In this control scheme, the actuation of the HP compressor- and turbine bypasses are regulated synchronously. The mapping that switches between 2-stage and 1-stage mode is a function of engine load and speed as was shown in Figure 4.10.
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Figure 5.1 shows the progress of turbocharger shaft speeds and diameters of the corresponding bypasses and wastegate. The synchronous actuation of the HP compressor- and turbine bypasses is evident up to around t=2 seconds, after which the compressor bypass continues to open past that of the turbine due to the former‟s larger full-open diameter. The closing of the bypass during the deceleration phase is also synchronous.
Figure 5.1: Turbocharger speeds and diameters of wastegate and bypasses; simultaneous bypass actuation; regulated wastegate before bypass opening
The progress of shaft speed shows that opening of the bypass valve is marked by simultaneously rapid deceleration of the HP turbocharger and acceleration of the LP unit. When both bypasses are completely open, the HP turbocharger essentially freewheels and continues to decelerate. Further investigation should be made on the test-bench to see whether the extremely rapid acceleration of the LP turbocharger during bypass opening in this mode, shown from 90000 to nearly 125000 RPM between t=1.7–1.8 sec. is also observed. If so, this represents a potential critical operating mode due to the large magnitude of shaft torque that would be transmitted.
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Figure 5.2: Manifold pressures and BMEP; simultaneous bypass actuation; regulated wastegate before bypass opening
Figure 5.3: Turbine powers; simultaneous bypass actuation; regulated wastegate before bypass opening
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Figure 5.2 shows the progress of intake- and exhaust manifold pressures as well as BMEP. The opening of the bypasses at t=~1.7 sec. is marked by a drastic reduction of exhaust backpressure. As seen by the blue curve, it falls from nearly 5 bar to under 2.5 bar nearly instantaneously, while the boost pressure has dropped approximately 0.6 bar during this same time window. The reduction of exhaust manifold pressure is a result of the switch from 2-stage to single-stage charging and the resulting higher total efficiency, which can be seen in a time plot of the sum of turbine powers in Figure 5.3. When the HP turbocharger is bypassed, the turbine power drops to negligible values (green curve). From that point on, all of the turbine work is performed by the LP turbocharger.
Figure 5.4: PMEP, BMEP, gross IMEP and exhaust manifold pressure; simultaneous bypass actuation; regulated wastegate before bypass opening
The large change in exhaust manifold pressure during switchover from 2-stage to single-stage charging mode has a significant effect on pumping mean effective pressure (PMEP). As Figure 5.4 shows, the PMEP changes by approximately 2 bar between t=1.7–1.8 sec. The large difference in PMEP results in a higher necessary gross IMEP before the switchover in order to maintain a given BMEP. This difference gives rise to the requirement for different boost pressure before- and after switching between 1- and 2-stage turbocharging modes in order to maintain the brake operating point. This observation justified the control of the wastegate by BMEP itself
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rather than boost pressure, since a mapped boost pressure at a given operating point before the switchover would result in different brake engine output afterwards as a result of the difference in PMEP. A solution to this while retaining control by boost pressure is to have separate maps of boost pressure in 1- and 2-stage modes. While this option has not been investigated, it is foreseeable that the dynamic transition between modes will be problematic for effective control.
Figure 5.5: Residual gas fraction as function of exhaust manifold pressure; simultaneous bypass actuation; regulated wastegate before bypass opening
The difference in exhaust manifold pressure also influences the trapped residual gas fraction. This is not so critical in a typical switch from 2-stage to 1-stage mode during acceleration, as the residual fraction drops in relation to exhaust manifold pressure as seen in Figure 5.5. However, as the residual fraction reaches a peak before the switchover, the ignition timing must be retarded in this vicinity to protect against knocking. Further development of the engine control unit can place a feedforward in the ignition timing control in order to re-advance the spark timing after the switchover under conditions of reduced residual fraction, which is less prone to knocking. This can act as an additional degree of freedom to regulate the engine output in order to improve smoothness, since the boost pressure inevitably falls during the changeover from 2-stage to 1-stage mode as the LP turbocharger must accelerate rapidly to a new operating point at higher pressure ratio in order to take over the compression
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work previously shared with the HP unit. The high residual gas fraction before the start of full-load acceleration is due to throttling at very low engine loads and a high degree of valve overlap. The high sensitivity of residual gas fraction to the exhaust manifold pressure is due to the fact that cam phaser angles were maintained fixed, rather than varying with load and speed. Therefore, the investigated timings optimized for a specific operating point and held constant result in a very high residual gas fraction during the constant speed portion of the drive at low-part loads.
Figures 5.6 to 5.8 show operating points plotted on compressor and turbine maps of both HP and LP stages. HP and LP compressor maps are illustrated separately but turbine maps are combined. Figure 5.6 shows that a larger trim or whole other larger impeller diameter could be selected to better fit the operating points in the region of best efficiency of the LP compressor. A larger impeller size will, however, increase the moment of inertia and impair transient response. The drop-off of operating points coincides with deactivation of the HP turbocharger during bypass opening.
Figure 5.6: HP compressor operating points; simultaneous bypass actuation; regulated wastegate before bypass opening
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Figure 5.7: LP compressor operating points; simultaneous bypass actuation; regulated wastegate before bypass opening
Figure 5.8: HP and LP turbine operating points; simultaneous bypass actuation; regulated wastegate before bypass opening
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67
Figure 5.7 shows that the switchover from 2-stage to 1-stage mode is marked by a brief incursion into the surge region. These incursions are very shortly lived, however. Each point represents an elapsed time of less than 0.08 seconds. However, further test bench investigations should still be performed in order to validate this observed phenomenon in the simulation mode and also to determine if long-term durability of the turbocharger could be compromised. The plot of operating points of the LP turbine in Figure 5.8 shows the switchover occurring with a rapid shift to a lower pressure ratio before rising again. This follows the progress of exhaust manifold pressure, as shown in Figure 5.5.
5.1.2
Delayed Opening of Compressor Bypass; LP Wastegate Also Regulated Before Bypass Opening
In this control strategy, the opening of the compressor bypass is delayed by fixed, constant time duration; in this case, a value of 0.5 sec. is investigated. The actuation of the compressor- and turbine bypasses are illustrated in Figure 5.9. As expected, the progress of turbocharger shaft speeds and boost pressure prior to the first bypass opening remains identical from the previously investigated case.
Figure 5.9: Turbocharger speeds and diameters of wastegate and bypasses; delayed opening of compressor bypass; regulated wastegate before bypass opening
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Figure 5.10: Manifold pressures and BMEP; delayed opening of compressor bypass; regulated wastegate before bypass opening
The delayed opening of the compressor bypass results in a significant overshoot in BMEP, as shown in Figure 5.10. This is because the opening of the turbine bypass before that of the compressor relieves the exhaust manifold pressure first, resulting in a drastic and almost instantaneous drop in PMEP. This occurs before being followed by a delayed response in boost pressure. Once the compressor bypass opens, the boost pressure then falls significantly when switching from 2-stage to 1-stage mode, because the supply of HP turbine work has been cut-off with the earlier opening of the turbine bypass, while the angular momentum of the HP turbocharger is dissipated in still compressing air in 2-stage mode during the delayed opening period of the compressor bypass. As a result, the drop in HP turbocharger shaft speed is much more drastic than in the previous case of synchronous opening of both bypasses.
Furthermore, the LP compressor is not able to accelerate to the new operating point in 1-stage mode sufficiently quickly to make up for the lost compression work performed in series. A drastic undershoot in engine BMEP results, which takes a substantial amount of time in order to recover back to the target value of 26 bar. The rise in LP shaft speed is more gradual than in the case of synchronous bypass
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opening. The lower rate of shaft RPM increase is therefore less of a concern than in the synchronous bypass opening case.
Figure 5.11: HP compressor operating points; delayed opening of compressor bypass; regulated wastegate before bypass opening
Figure 5.12: LP compressor operating points; delayed opening of compressor bypass; regulated wastegate before bypass opening
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Figure 5.13: HP and LP turbine operating points; delayed opening of compressor bypass; regulated wastegate before bypass opening
Figures 5.11 and 5.12 show the transient progress of HP- and LP compressor operating points on respective maps. The switchover from 2-stage to 1-stage mode is differentiated here by the mass flow rate dropping off to the left of the trace during the acceleration phase instead of to the right with synchronous opening shown in Figure 5.6 because the HP turbine work is being withdrawn earlier via the HP turbine bypass than the compressor bypass.
On the plot of LP compressor operating points, there is a lesser degree of incursion into the surge region during switchover. This is because the delayed opening of the compressor bypass allows more time for pressures on either side to equalise on their own, due to the decline of HP compressor pressure ratio with the advanced opening of the turbine bypass.
Results
5.1.3
71
Delayed Opening of Compressor Bypass; LP Wastegate Remains Closed Before Bypass Switching
A strategy is investigated that keeps the wastegate closed before opening of the bypasses. This in fact follows the control strategy implemented in the demonstrator vehicle and described by Figure 4.14 in Section 4.3.1. This is performed on the case where the compressor bypass opening is delayed by 0.5 seconds as discussed previously. In theory, keeping the wastegate closed before switchover should improve transient response, since the more of the available energy of exhaust gas works upon the LP turbine rather than being unutilised.
Indeed, compared to Figure 5.9 wherein the wastegate begins to open earlier, Figure 5.14 shows that in this case the shaft speed of the LP turbocharger is higher in the first moments after the HP turbine bypass opens. At t=1.7 sec., the LP shaft speed is approximately 125000 RPM compared to 100000 RPM. This advantage of higher LP shaft speed continues until complete opening of the HP compressor bypass.
Figure 5.14: Turbocharger speeds and diameters of wastegate and bypasses; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening
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Figure 5.15: Manifold pressures and BMEP; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening
As a result of the higher LP turbocharger RPM, there is a smaller undershoot of boost pressure and BMEP, as illustrated in Figure 5.15 (compare with Figure 5.10). The BMEP is also able to recover more quickly to the target value.
Figure 5.17 shows that the switchover from 2-stage to 1-stage mode is marked by only a slight incursion into the surge region immediately followed by some instability of the mass flow rate, as evidenced by a near-horizontal oscillation before continuing onwards. This may be explained by the fact that prior to the switchover, there is a pressure difference at both sides of the compressor bypass. The rapid opening of the compressor bypass results in mass transfer in order to equalise the pressures. This equalisation occurs with oscillations in mass flow rates and pressure ratios across the compressor, which is laid out in parallel to the bypass.
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Figure 5.16: HP compressor operating points; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening
Figure 5.17: LP compressor operating points; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening
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Figure 5.18: HP and LP turbine operating points; delayed opening of compressor bypass; wastegate remains closed before HP turbine bypass opening
5.1.4
Compressor Bypass Remains Closed Throughout Acceleration; LP Wastegate Remains Closed Before HP Turbine Bypass Opening
This investigated case was to study the behaviour of the turbocharger operating points and dynamic characteristics when the compressor bypass remains closed. This is a subject of interest, since acceptable operation in this mode will make a compelling case to eliminate the need for the compressor bypass altogether, thus simplifying control, making calibration easier as well as reducing parts count and cost.
Figure 5.21 shows the progress of turbocharger shaft speed. Once the turbine bypass is completely open, the shaft speed of the HP turbocharger rises again. This is because the full charge air mass flow is not bypassed but rather continues to flow through HP the compressor. In effect, the compressor acts as a turbine, with the through-flowing air driving the turbocharger. Furthermore, there exists a pressure drop from the inlet to the outlet, that is, the pressure ratio:
C , HP
p2, HP p1, HP
1
Equation 5.1
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Compressor maps are not typically published for pressure ratios below 1. GT-SUITE extrapolates speed curves from available data to fill the missing operating points. The behaviour of the compressor at C 1 has not yet been extensively investigated and was not observed during bench testing; therefore simulated predictions using extrapolated maps need to be further validated.
Figure 5.19: Turbocharger speeds and diameters of wastegate and bypasses; compressor bypass always closed; wastegate remains closed before HP turbine bypass opening
Analysing the progress of manifold pressures and BMEP in Figure 5.20 shows that the boost pressure and BMEP suffer from severe undershoot during opening of the turbine bypass. The reason is that when the HP stage is deactivated while the compressor bypass remains closed, as explained earlier, the HP compressor acts as a restriction. The LP compressor therefore must traverse a larger change of operating point from 2-stage compression to a higher pressure ratio in single-stage mode in order to overcome the restriction across the HP compressor and maintain a given intake manifold pressure. This is evident in the plot of operating points on the LP compressor map shown in Figure 5.22, which shows that the pressure ratio reaches a higher peak value than in all previously investigated cases.
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Figure 5.20: Manifold pressures and BMEP; compressor bypass always closed; wastegate remains closed before HP turbine bypass opening
Figure 5.21: HP compressor operating points; compressor bypass remains closed; wastegate remains closed before HP turbine bypass opening
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Figure 5.22: LP compressor operating points; compressor bypass remains closed; wastegate remains closed before HP turbine bypass opening
Figure 5.23: HP and LP turbine operating points; compressor bypass remains closed; wastegate remains closed before HP turbine bypass opening
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Figure 5.24 shows that unlike the previous two cases with delayed opening of the compressor bypass, there is no associated equalisation of mass flow. All investigations thus far have concluded that a different LP compressor optimised for a somewhat lower mass flow would be a better match for this engine and demanded operating range. In fact, it is possible to select a compressor that simultaneous gives better efficiency over a wider range of operating conditions at full-load, whilst clearing the surge line during switchover to single-stage mode.
Further development of the wastegate control strategy and implementation of the aforementioned spark timing advance immediately after HP turbine bypass opening could mitigate the torque over- and undershoots, making eliminating the HP compressor bypass a viable option. The higher shaft speed of the HP turbocharger when bypassed could be beneficial in normal driving that is not characterised solely by monotonous acceleration but by alternating acceleration and deceleration. However, this phenomenon has only observed during simulations with GT-SUITE and should be further validated on the test bench . 5.1.5
Simultaneous
Regulated
Opening
of
Compressor-
And
Turbine
Bypasses; LP Wastegate Remains Closed Before Bypass Switching The final investigated case considers once again a synchronous actuation of the HP compressor- and turbine bypasses, but with the LP wastegate closed before the switchover from 2-stage to single-stage mode.
Figure 5.26 shows the usual progress of shaft speeds and behaviour of the wastegate and both HP bypasses. The dynamic torque generation was the smoothest of all investigated cases, as shown in Figure 5.27. There is little change in BMEP during the switchover from 2-stage to single-stage operation.
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Figure 5.24: Turbocharger speeds and diameters of wastegate and bypasses; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening
Compared with Figure 5.1, the LP turbocharger reaches a higher RPM during the opening of the bypasses. Figure 5.24 shows that at about t=1.7 sec., the LP turbocharger shaft speed is about 145000 compared to 130000 RPM, while the LP shaft speed is little changed. The rate at which the LP turbine accelerates during switchover requires more investigation. One possible measure is to actuate the bypasses at a slower rate, instead of going to full-open in 0.3 seconds. This should be weighed with any potential impact on smooth torque generation.
According to the simulation results, this strategy results in the smoothest progress of engine torque. The BMEP rises to the target of 26 bar quickly and remains very steady.
As is to be expected the traces of compressor and turbine operating points are very similar to those obtained in Section 5.1.1
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Figure 5.25: Manifold pressures and BMEP; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening
Figure 5.26: HP compressor operating points; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening
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Figure 5.27: LP compressor operating points; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening
Figure 5.28: HP and LP turbine operating points; simultaneous bypass actuation; wastegate remains closed before HP turbine bypass opening
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Figure 5.29: Summary plot of engine speeds of all investigated control strategies
Figure 5.29 plots the progress of engine speed for the 5 investigated control strategies during full-load acceleration from 20 km/h in second gear, while Figure 5.30 shows vehicle speed. All cases resulted in similar performance, but the best was achieved by simultaneous bypass actuation and maintaining the wastegate closed before the switch from 2-stage to single-stage mode. This was achieved though having the smoothest torque development with little undershoots of engine output.
Summarising the results of this chapter, 5 different strategies of controlling the various bypasses and wastegate have been investigated. Each gives different behaviour in regards to operating point and the smoothness of torque generation. All of the investigated strategies resulted in vehicle performance that was very similar; only approximately 0.2 seconds separated the quickest and slowest strategy in simulated full-load acceleration from 20 – 70 km/h.
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Figure 5.30: Summary plot of vehicle speeds of all investigated control strategies
Figure 5.31: Summary plot of vehicle acceleration of all investigated control strategies
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However, plots of BMEP – or vehicle acceleration as shown in Figure 5.31 – show large differences in dynamic behaviour. The relatively large variations in BMEP demonstrated by some concepts may be noticeable and subjectively uncomfortable. As a whole, however, the strategy of synchronised actuation of both the HP compressor and HP turbine bypasses gives the smoothest torque development of all investigated cases, particularly when the LP wastegate remains closed before switchover.
5.2 Stationary Simulation A stationary load step investigation involves allowing the engine to first settle at the initial operating point and then applying wide-open throttle while maintaining constant speed. This type of test can be performed on an engine test bench using an eddycurrent dynamometer that holds the engine speed fixed with different input load.
5.2.1
Load Step 2000 RPM / 2 bar BMEP
Figure 5.32 shows the result of the load step for the GT² in comparison with the SGT. At time t=0, the throttle is opened to the 100% position. Plots of intake manifold pressure, IMEP and turbocharger shaft speeds are shown. The 2-stage charging concept results in a substantially higher attained IMEP and intake manifold pressure. More importantly, the respective rates of increase are also higher. The shaft speed of the HP-turbocharger stage of the GT² is able to accelerate at a faster rate than the LP stage and the single turbocharger of the SGT. This validates the benefit of having a low moment of inertia of the small HP turbocharger in reducing turbocharger lag.
The GT² reaches the rated IMEP (corresponding to 26 bar BMEP) in 1.6 seconds, compared to 3.8 seconds of the SGT to reach 22.3 bar BMEP. The time for intake manifold pressure to build up follows the trend established for the IMEP; the GT² reaches 2.4 bar in 1.4 seconds, while the SGT reaches 1.9 bar in 3.3 seconds.
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Figure 5.32: Comparison of constant speed load-step performance (red – GT²; blue – SGT)
Characteristic of all load step investigations, whether stationary or transient, is the observation that the intake manifold pressure and IMEP increase rapidly in the first fractions of a second upon sudden opening of the throttle, up to a certain point where the manifold pressure rises from the throttled condition until stabilizing to a balanced point at the inlet and outlet of the throttle near ambient. The IMEP at this point is at about the magnitude of a naturally-aspirated engine at the same speed and wideopen throttle position. The turbocharger(s) is/are accelerating but due to rotating inertia and the finite time of filling the volume between the compressor outlet and intake manifold, is not contributing to any substantial build-up of boost pressure in these first moments. As the turbocharger continues to accelerate, the boost pressure and engine output increase correspondingly until regulated at the full-load line by means of a wastegate, for example. Some overshoot and oscillation of boost pressure and engine output around the target are normal phenomena, but should be minimised not only for comfort reasons but also to protect against instances where overshoots could result in knocking and engine damage.
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Figure 5.33: Compressor operating points for SGT; load step from 2 bar BMEP / 2000 RPM to full-load at constant engine speed.
The large different in intake manifold pressure during the constant running phase of the simulation has been found to be due to very a very high residual gas fraction at low part load. This has been traced to a non-optimal cam phaser timing that results in late intake valve closing, which causes some charge air to be pushed out of the cylinder . The lost trapped mass must be made up for through higher intake manifold pressures.
Due to the constant engine speed of this investigated simulation, the aspiration curve essentially follows a relationship with rising boost pressure, as shown in Figure 5.33. While this is not a common operating regime in a real vehicle, care, should be taken in order to ensure a sufficient surge margin, especially where higher altitudes or high compressor inlet temperatures are encountered.
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Figure 5.34 shows the turbine operating points bending down from the upper boundary of the map, at the wastegate opens.
Figure 5.34: Turbine operating points for SGT; load step from 2 bar BMEP / 2000 RPM to full-load at constant engine speed.
Figure 5.35: HP and LP compressor operating points for GT²; load step from 2 bar
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BMEP / 2000 RPM to full-load at constant engine speed.
Figure 5.36: HP and LP turbine operating points for GT²; load step from 2 bar BMEP / 2000 RPM initial condition to full-load at constant engine speed.
Figures 5.35 and 5.36 show the traces of operating points on the compressor and turbine maps, respectively, for the GT². HP- and LP maps for both compressors and turbines are shown on the same graphs on common axes, respectively.
5.3 Transient Simulations The previous section showed results of a load step simulation at a fixed RPM point with only the engine model. The results of this chapter and those going forward are with a coupled engine/vehicle model. These simulations therefore show the behaviour of the engine integrated into the chassis and can be readily validated with road-test measurements.
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89
Coupled Full-Load Acceleration 25 km/h in 3rd Gear
Road test measurements have been performed in third gear at an initial speed of 25 km/h corresponding to an engine speed of 1220 RPM. These initial conditions have been re-created in the coupled model in GT-SUITE in order to calibrate and validate the simulation with measured real-world results. The measured data are obtained by sensors mounted into the vehicle during testing and stored in a mobile data acquisition system. The collected wastegate and bypass signals are inputted as imposed values into the GT-SUITE model under the same operating regime (full-load acceleration at stated initial vehicle speeds and selected gear). These signals are then scaled with a multiplication factor until a suitable correlation of measurement data with the simulation is achieved, since the measured data represent linear transducer signal values for measurement of the bypasses and PWM (pulse-width modulation) duty cycles for the LP wastegate. Therefore, these signals do not correspond exactly on a 1:1 basis with actuated values in reality and in the model.
Figure 5.37: Engine speeds; acceleration from 25 km/h in 3rd gear
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Figure 5.38: Turbocharger shaft speeds; acceleration from 25 km/h in 3rd gear
Figure 5.37 shows plots of engine speed (solid – simulated; dashed – measured). In this scenario, the vehicle is driven at an initial constant speed of 25 km/h in third gear. Corresponding with t=0 sec., the driver invokes full-throttle acceleration in the same gear. The correlation between simulation and measurement values is very good: about 0.2 seconds separate the two results at 90 km/h corresponding to approx. 3000 RPM.
Figure 5.38 shows the calibration of turbocharger shaft speed. The correlation between measurement and simulation results is also good. The correlation of intake manifold pressure shows some deviations, particularly during the constant driving phase and also in the first moments of the full-load acceleration run. This is again caused the extrapolation of the compressor map done internally in GT-SUITE. At some operating points in this low part-load operating regime, the compressor could be working at pressure ratios very close to- or even less than unity. The extrapolation of efficiencies in this region yields significant error, which affect the calculated boost pressure based on available turbine work.
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Figure 5.39: Intake manifold pressure; acceleration from 25 km/h in 3rd gear
Figure 5.40: Compressor operating points; acceleration from 25 km/h in 3rd gear
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Figure 5.42 shows the progress of compressor operating points, with maps for the HP- and LP stages superimposed on common axes.
Figure 5.41: Turbine operating points; acceleration from 25 km/h in 3rd gear Figure 5.43 shows the progress of turbine operating points, with maps for the HPand LP stages superimposed on common axes. Analysis of the measured wastegate signals revealed that the high exhaust manifold pressures encountered while driving were sufficient to overcome the spring force that keeps the wastegate closed. As a result, the true potential performance and progress of turbocharger operating points cannot be ascertained from these results.
5.3.2
Coupled Full-Load Acceleration 20 km/h in 2nd Gear
In order to benchmark the transient performance of the GT² concept, the acceleration performance from 20 km/h in second gear is compared with a GT-DRIVE simulation of a naturally-aspirated 3.5 L engine that has been “virtually-integrated” into the same Ford Focus chassis with identical gear ratios and other boundary conditions. The GT² simulations are the same as in Sub-section 5.1.5, since it gave the best performance and smoothness.
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Figure 5.42: Full-load torque and power; 3.5L naturally-aspirated benchmark engine
The full-load torque curve of the benchmarked naturally-aspirated engine, shown in Figure 5.42, is taken from a representative vehicle in this class, a Mercedes-Benz CLS350. The GT-DRIVE model calculates the vehicle acceleration from the quotient of the excess force and inertial mass. This is calculated after matching demanded road resistance torque referred back to the engine crankshaft to mapped brake operating points without any slip in the drivetrain, so that specific details about the throttle response programming as well as vehicle application and calibration are neglected.
Figure 5.43 shows a time plot of both benchmarked cases. The progress of vehicle speed and acceleration are illustrated. What is not surprising is that the shapes of the acceleration curves follow that of the progress of developed brake torque of each engine. The naturally-aspirated engine reaches the full-load torque curve after about 0.5 seconds. The GT² by comparison reaches its rated torque curve after 1.7 seconds due to turbo lag. This is considered excellent for a turbocharged engine, in light of the fact that this scenario of accelerating from 20 km/h in second gear represents a worst-case, due to the very low initial engine speed of about 1390 RPM and the highly dynamic acceleration in second gear. In spite of this worst-case
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scenario, the acceleration to 70 km/h is completed in 3.3 seconds for the GT² compared to 2.7 seconds in the 3.5 L naturally-aspirated example.
Figure 5.43: Comparison of vehicle speed and acceleration, second gear from 20 km/h
Since the GT² acceleration plot and setup is the same as that detailed in Section 5.1.5, compressor- and turbine operating point maps are not shown again.
5.3.3
Coupled Full-Load Acceleration 30 km/h in 2nd Gear
In order to benchmark the performance of the GT² concept with measurement results from a database of competitor vehicles, acceleration in second gear from an initial speed of 30 km/h is simulated. Figure 5.44 compares once again the GT² Ford Focus with the aforementioned virtual counterpart with the 3.5L naturally-aspirated engine. The higher initial engine speed reduces the impact of turbo lag and results in shorter time to reach the full-load torque curve, ca. 1.4 seconds. The GT² reaches 80 km/h in 3.1 seconds compared to the 3.5L NA counterpart‟s 2.7 seconds.
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Figure 5.44: Comparison of vehicle speed and acceleration, second gear from 30 km/h
Figure 5.45: Comparison of full-load acceleration of FEV GT² from 30 km/h in second gear vs. competitors
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Figure 5.45 compares acceleration of the GT²-equipped Ford Focus with 2 representative competitors in the same segment. The Audi A3 employs a 2.0L gasoline direct-injected engine with single-stage turbocharging. The platform-twin to the Audi A3, the Golf TSI, incorporates a 1.4L engine with a combination of an exhaust gas turbocharger and mechanical supercharger. The figure shows two notable features. The first is that the peak acceleration magnitude of the GT² Ford Focus is substantially higher than both competitors on account of its much higher peak torque value of 375 Nm. This high magnitude of acceleration is maintained for a comparatively long period of time, the function of a constant full-load torque curve over a wide range of engine speed. The second feature is that the rate of acceleration, that is, the slope the acceleration curve with respect to time, is substantially greater than that of the single-stage turbocharged Audi A3 and approaches that of the twin-charged Golf TSI. Both illustrated quantitative factors combine to give the GT² a subjectively very high fun-to-drive factor.
Figure 5.46: Compressor operating points; acceleration from 30 km/h in 2nd gear
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Figure 5.47: Turbine operating points; acceleration from 30 km/h in 2nd gear
Figure 5.46 shows the progress of compressor operating points, with maps for the HP- and LP stages superimposed on common axes. Figure 5.47 shows the progress of turbine operating points, with maps for the HP- and LP stages superimposed on common axes. Adopting the same control strategy as that for the acceleration from 20 km/h in second gear, the transient plots of compressor- and turbine operating points result. The LP compressor operates at a pressure ratio of nearly 3.2; clearly this will require substantial mixture enrichment in order to maintain target turbine-inlet temperatures. The LP compressor, which is rather too large at lower engine speeds and power outputs for best efficiency, is well laid out for this requirement. However, the HP compressor, which was already seen to be too small, is clearly much undersized in this acceleration scenario.
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6 Discussion In this chapter, the results from the previous chapter will be briefly summarised in the context of the objectives for 2-stage turbocharging. It will be followed by an outlook of potentials for further development. Some concluding statements will end this chapter and this thesis report.
6.1 Summary This thesis has detailed the development of control strategies for realising 2-stage turbocharging in Otto engines. This has been achieved via simulations performed upon coupled and calibrated models of the detailed engine and vehicle/chassis in GT-SUITE. Investigations have been undertaken in order to study the effects of different control strategies of the various bypasses and wastegate used in the 2stage concept. The simulation studies suggest that a simultaneous opening of the HP turbine- and compressor bypasses would give the smoothest transition in engine torque during the switchover from 2-stage to single-stage charging mode compared to other investigated strategies. Elimination of the HP compressor bypass – simulated by simply leaving it always closed – does not give as smooth transition than the aforementioned simultaneous actuation of both bypasses. The simulation in this case shows that the HP turbocharger, being driven by the air flowing through the respective compressor, does not reach critical operating speeds when the engine is accelerated to its rated speed and power points. Moreover, the HP turbocharger – although idling but not bypassed on the compressor side – acts as pressure drop. Therefore, in single-stage mode, the LP compressor must operate at a higher map pressure ratio in order to overcome the restriction on the HP side to achieve the same manifold pressure. Further optimization of the control system – combined with adding a degree-of-freedom for advancing the spark timing during switchover under conditions of reduced knocking propensity – could minimise the torque fluctuations that could be noticeable by the driver to indiscernible levels.
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A stationary load-step from an initial 2 bar BMEP / 2000 RPM operating point shows that the 2-stage GT² concept reaches the target boost pressure and rated full-load line much quicker than the single-stage SGT counterpart. The GT² reaches the rated IMEP in 1.6 seconds, compared to 3.8 seconds of the SGT, although the output of the GT² is higher at 26 bar BMEP versus 22.3 bar of the SGT.
Investigations of full-load transient acceleration from various starting vehicle speeds and selected gears also show excellent performance of the GT² concept, especially when compared against simulation results of a 3.5L naturally-aspirated engine and road-test measurements of 2 benchmark competitors. Simulations of acceleration from 30-80 km/h in second gear show that the 1.8L GT² achieves an elapsed time of 3.1 seconds; the 3.5L counterpart completes the same run in 2.7 seconds. In the same test, the GT² achieves rated torque after 1.4 seconds, which is also excellent. The GT² has a much higher peak acceleration value than that of a production singlestage turbocharged vehicle in the same class (approx. 5.5 vs. 3 m/s²). The acceleration of the GT² remains competitive to another production vehicle in the same class equipped with an engine that combines a turbo- and supercharger.
6.2 Outlook The course of this thesis has identified some areas for further research and development for two-stage turbocharging, particularly its novel application in Otto engines. The limited time for the project precluded deeper investigation in certain areas than desired, particularly in the objective for transient simulation of a coupled engine/vehicle model in GT-SUITE as a supporting tool for vehicle application. Nonetheless, this thesis has revealed insight into the operational behaviour of the turbochargers and engine performance, particularly the effect of control strategies of the various bypasses and wastegate during the switchover from 2-stage to singlestage mode during acceleration. The insight and predictions from simulation are valuable to determine what operational conditions should be avoided in the vehicle calibration; for example determining the threshold in the engine operating map for switchover between 2- and single-stage mode, while leaving sufficient margin from maximum turbocharger shaft speed. While project timing resulted in the boost control system being developed separately, the approaches proved to be broadly similar between the model in GT-SUITE and the in-vehicle implementation.
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Future work can integrate the simulation and calibration more closely by extending the current coupled engine-vehicle model with the control system modelled in Simulink®. This Simulink® model can then be implemented as a functional softwarein-the-loop engine control unit with dSpace. This can accelerate and optimise the development of the control system by building only one model that will work in both GT-SUITE and in a hardware control unit. Predictions from the simulations can be quickly validated on the test bench and road-testing. This can be combined with more detailed modelling of the physical wastegate and bypass valves. Further optimization of the engine can be realised through re-matching of the turbochargers and improved valve timing control. As a next step, the GT² engine concept should be integrated into a larger vehicle to better demonstrate its fuel saving credentials through downsizing rather than as a high-performance demonstrator.
6.3 Conclusions Two-stage turbocharging of Otto engines has been shown in this thesis to nearly preserve the excellent transient performance and fun-to-drive factor of a large naturally-aspirated counterpart while leveraging the reduced driving cycle fuel consumption that can be realised by downsized engine displacement.
Simulations of coupled engine- and vehicle models in GT-SUITE have been calibrated with good correlation with road-test measurement data. At a downsizing factor of nearly 50% compared with a naturally-aspirated baseline engine, the acceleration from 30-80 km/h in second gear is 0.4 seconds slower at 3.1 instead of 2.7 seconds.
By demonstrating far better transient response than conventional turbocharged engines, while offering competitive performance compared to a large-displacement naturally-aspirated one as well as another downsized variant with both a turbo- and supercharger, this thesis has shown the viability of 2-stage turbocharging realised in the GT² concept as an effective means to bridge the compromise of transient performance in downsized Otto engines.
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Bibliography
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GT-SUITE Online Help Navigator, content entry “PIDController”, Version 7.0, 2009
