Reciprocating Internal Combustion Engines

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1 Part 2: Turbochargers, Engine Performance Metrics Reciprocating Internal Combustion Engines Prof. Rolf D. Reitz Engine Research Center University of Wisconsin-Madison 2014 Princeton-CEFRC Summer School on Combustion Course Length: 15 hrs (Mon.- Fri., June 23 27, 2014) Copyright 2014 by Rolf D. Reitz. This material is not to be sold, reproduced or distributed without prior written permission of the owner, Rolf D. Reitz. 1 CEFRC1-2, 2014

2 Part 2: Turbochargers, Engine Performance Metrics Short course outine: Engine fundamentals and performance metrics, computer modeling supported by in-depth understanding of fundamental engine processes and detailed experiments in engine design optimization. Day 1 (Engine fundamentals) Part 1: IC Engine Review, 0, 1 and 3-D modeling Part 2: Turbochargers, Engine Performance Metrics Day 2 (Combustion Modeling) Part 3: Chemical Kinetics, HCCI & SI Combustion Part 4: Heat transfer, NOx and Soot Emissions Day 3 (Spray Modeling) Part 5: Atomization, Drop Breakup/Coalescence Part 6: Drop Drag/Wall Impinge/Vaporization/Sprays Day 4 (Engine Optimization) Part 7: Diesel combustion and SI knock modeling Part 8: Optimization and Low Temperature Combustion Day 5 (Applications and the Future) Part 9: Fuels, After-treatment and Controls Part 10: Vehicle Applications, Future of IC Engines 2 CEFRC1-2, 2014

3 Part 2: Turbochargers, Engine Performance Metrics Turbocharging Pulse-driven turbine was invented and patented in 1925 by Büchi to increase the amount of air inducted into the engine. - Increased engine power more than offsets losses due to increased back pressure - Need to deal with turbocharger lag Improved 3 CEFRC1-2, 2014

4 Turbocharging Part 2: Turbochargers, Engine Performance Metrics Purpose of turbocharging or supercharging is to increase inlet air density, - increase amount of air in the cylinder. Mechanical supercharging - driven directly by power from engine. Turbocharger - connected compressor/turbine - energy in exhaust used to drive turbine. Supercharging necessary in two-strokes for effective scavenging: - intake P > exhaust P - crankcase used as a pump Some engines combine engine-driven and mechanical (e.g., in two-stage configuration). Intercooler after compressor - controls combustion air temperature. 4 CEFRC1-2, 2014

5 Turbocharging Part 2: Turbochargers, Engine Performance Metrics Energy in exhaust is used to drive turbine which drives compressor Wastegate used to by-pass turbine Charge air cooling after compressor further increases air density - more air for combustion 5 CEFRC1-2, 2014

6 Part 2: Turbochargers, Engine Performance Metrics Regulated two-stage turbocharger Duplicated Configuration per Cylinder Bank LP stage Turbo-Charger with Bypass Compressor Bypass HP stage Turbo charger Charge Air Cooler Regulating valve EGR Cooler EGR Valve GT-Power R2S Turbo Circuit EGR Valve EGR Cooler HP TURBINE Compressor Bypass Charge Air Cooler Compressor Bypass Regulating valve HP stage Turbo charger LP stage Turbo-Charger with Bypass Regulating Valve LP TURBINE LP Stage Bypass 6 CEFRC1-2, 2014

7 Part 2: Turbochargers, Engine Performance Metrics Intercooler for IVC temperature control Isentropic P P IVC V V IVC ln P ln T Q Reduced Peak Temp (NOx) Improved phasing T T IVC V V IVC ( 1) Pressure /time of ignition Boost Compressor T ign Q TDC IVC TDC ln V Boost explains 20% of the improved fuel efficiency of diesel vs. SI 7 CEFRC1-2, 2014 IVC ln V

8 Part 2: Turbochargers, Engine Performance Metrics Automotive compressor Centrifugal compressor typically used in automotive applications Provides high mass flow rate at relatively low pressure ratio ~ 3.5 Rotates at high angular speeds - direct coupled with exhaust-driven turbine - less suited for mechanical supercharging Consists of: stationary inlet casing, rotating bladed impeller, stationary diffuser (w or w/o vanes) collector - connects to intake system 8 CEFRC1-2, 2014

9 Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990 Compressible flow A review Gibbs Energy Tds dh dp / dh VdV Euler dp VdV Area-velocity relations for M<1 for M>1 AV Const da A 2 da (1 M ) 2 dp A d da dv A V 2 ( M 1) V dv V 0 Subsonic nozzle Subsonic diffuser Supersonic diffuser Supersonic nozzle da<0 da >0 da <0 da >0 from AV dv>0 dv <0 dv <0 dv >0 from Euler dp<0 dp >0 dp >0 dp <0 kinetic energy pressure recovery kinetic energy Traffic flow behaves like a supersonic flow! 9 CEFRC1-2, 2014

10 Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990 Model passages as compressible flow in converging-diverging nozzles P V m AV A RT RT c P AM ( P / P ) /( T / T ) RT0 1/ 2 P 0 A* Minimum area point With M=1: Fliegner s formula Choked flow, M=1 mm 1 2 2( 1) 1 ( ) 0 1 RT0 P A * 1 A*/A Subsonic Supersonic Area Mach number relations A * A 1 2 ( 1) M A 1 1 ( P ) 1 ( P ) * A P 0 P0 2 (1 M ) 1 2( 1) 1 1 1/ solutions for same area 1 reservoir throat exit P/P M 10 CEFRC1-2, 2014

11 Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990 Isentropic nozzle flows T T M P0 P (1 M ) 2 Ex. Flow past throttle plate P 0 y P 1 0 P 0 1 P=P b Choked flow for P 2 < 53.5 kpa = 40.1cmHg reservoir ambient WOT Choked P/P 0 1 P b m M=1 y Manifold pressure, P 1 cmhg x 11 CEFRC1-2, 2014

12 Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990 Application to turbomachinery Fliegner s Formula: mm 1 2 2( 1) 1 ( ) 0 1 RT0 P A * Variable Geometry Compressor/ turbine performance map Increased speed Corrected mass flow rate A measure of effective flow area m T 0 ref P / P / T ref 0 Reduced flow passage area Choked flow 1.0 1/0.528=1.89 P 0 /P Total/static pressure ratio 12 CEFRC1-2, 2014

13 Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988 T Compressor ( T c ( T outisen out Tin) T ) in P 0 3 P 3 P 2 = P out Heywood, Fig Air at stagnation state 0,in accelerates to inlet pressure, P 1, and velocity V 1. V 1 2 / 2 c P P 0 P 1 = P 0,in Compression in impeller passages increases pressure to P 2, and velocity V 2. Diffuser between states 2 and out, recovers air kinetic energy at exit of impeller producing pressure rise to, P out and low velocity V out Note: use exit static pressure and inlet total pressure, because kinetic energy of gas leaving compressor is usually not recovered S W m h h W c c a out in a 1 m a a cp T a in p out 1 c p 0, in 13 CEFRC1-2, 2014

14 Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988 Compressor maps Work transfer to gas occurs in impeller via change in gas angular momentum in rotating blade passage Surge limit line reduced mass flow due to periodic flow reversal/reattachment in passage boundary layers. Unstable flow can lead to damage Pressure ratio evaluated using total-to-static pressures since exit flow kinetic energy is not recovered Speed/pressure limit line Non-dimensionalize blade tip speed (~ND) by speed of sound At high air flow rate, operation is limited by choking at the minimum area point within compressor Supersonic flow Shock wave Heywood, Fig CEFRC1-2, 2014

15 Part 2: Turbochargers, Engine Performance Metrics Serrano, 2007 Compressor maps GM 1.9L diesel engine Pressure Ratio (t/t) Efficiency (T/T) Corrected Air Flow (kg/s) Corrected Air Flow (kg/s) CEFRC1-2, 2014

16 Part 2: Turbochargers, Engine Performance Metrics Reitz, 2007 Automotive turbines Naturally aspirated: P intake =P exhst =P atm ( ) Boosted operation: Negative pumping work: P 7 <P 1 but hurts scavenging P 3 4 W m ( h h ) W t g in 0, out m c T t g P in t P 1 P 0, out in 1 g g Compressor P intake P exhst Expansion 5 Compression 1 7 Blowdown Available work (area 5-6-7) 6 6 Turbine 6 P amb TDC BDC P-V diagram showing available exhaust energy - turbocharging, turbocompounding, bottoming cycles and thermoelectric generators further utilize this available energy V 16 CEFRC1-2, 2014

17 Turbochargers Part 2: Turbochargers, Engine Performance Metrics Radial flow automotive; axial flow locomotive, marine T V 1 2 / 2 c P P 0 = P 0,in P 1 P 2 P 03 m N corrected corrected m g p T N T3 T T p 0 0 out P 3 = P out t ( T ( T out outisen Tin) T in ) S 17 CEFRC1-2, 2014

18 Part 2: Turbochargers, Engine Performance Metrics Compressor selection To select compressor, first determine engine breathing lines. The mass flow rate of air through engine for a given pressure ratio is: = IMP = PR * atmospheric pressure (no losses) = IMT = Roughly constant for given Speed 18 CEFRC1-2, 2014

19 Compressor Pressure Ratio Part 2: Turbochargers, Engine Performance Metrics Engine breathing lines 3.8 Engine Breathing Lines 1.4L Diesel, Air-to-Air AfterCooled, Turbocharged Torque Peak (1700rpm) Trq Peak Operating Pnt Rated (2300rpm) Rated Operating Pnt Intake Mass Flow Rate (lb/min) Parameter Torque Peak Rated Units Horsepower hp BSFC lb/hp-hr A/F none 19 CEFRC1-2, 2014

20 a a g g p p m m T Cp T Cp p p mech c t air fuel a g W t = W c Heywood, CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics..

21 Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988 Ideal engine efficiency Otto cycle Maximum possible closed-cycle efficiency ( ideal efficiency ) State (1) to (2) isentropic (i.e., adiabatic and reversible) compression from max (V1) to min cylinder volume (V2) Compression ratio rc = V1/V2. State (2) to (3) adiabatic and isochoric (constant volume) combustion, State (3) to (4) isentropic expansion. State (4) to (1) exhaust process - available energy is rejected - can be converted to mechanical or electrical work: 21 CEFRC1-2, 2014

22 Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988 Ideal engine efficiency Otto cycle Efficiency = net work / energy supplied Otto [( T 3 T 4) ( T 2 T1)]/( T 3 T 2) 1( T 4 T1) /( T 3 T 2) However, T 2 3 W expansion T / T ( V / V ) r ( V / V ) T / T c W compression = / r 1 1 c s r c 22 CEFRC1-2, 2014

23 Part 2: Turbochargers, Engine Performance Metrics η ideal Function of only two variables, compression ratio (r c ) and ratio of specific heats (γ) Increasing r c increases operating volume for compression and expansion Increasing γ increases pressure rise during combustion and increases work extraction during expansion stroke. Both effects result in an increase in net system work for a given energy release and thereby increase engine efficiency. Actual closed-cycle efficiencies to deviate from ideal: 1.) Assumption of isochoric (constant volume) combustion: Finite duration combustion in realistic engines. Kinetically controlled combustion has shorter combustion duration than diesel or SI - duration limited by mechanical constraints, high pressure rise rates with audible engine noise and high mechanical stresses 2.) Assumption of calorically perfect fluid: Specific heats decrease with increasing gas temperature; species conversion during combustion causes γ to decrease 3.) Adiabatic assumption: Large temperature gradient near walls results in energy being lost to heat transfer rather than being converted to crank work 23 CEFRC1-2, 2014

24 Part 2: Turbochargers, Engine Performance Metrics Herold, 2011 Other assumptions: In engine system models, compressors, supercharger, turbines modeled with constant isentropic efficiency instead of using performance map. - typically, compressors, superchargers, and fixed geometry turbines have isentropic efficiencies of 0.7. VGT has isentropic efficiency of Charge coolers - intercooler, aftercooler, and EGR cooler modeled with zero pressure drop, a fixed effectiveness of 0.9, constant coolant temperature of 350 K. 24 CEFRC1-2, 2014

25 Part 2: Turbochargers, Engine Performance Metrics Herold, 2011 Zero-dimensional closed-cycle analysis: Combustion represented as energy addition to closed system Fuel injection mass addition from user-specified start of injection crank angle (θ SOI ) and injection duration (Δθ inj ). Pressure and mass integrated over the closed portion of cycle with specified initial conditions at IVC of pressure (p 0 ), temperature (T 0 ), and composition (x n,0 for all species considered - N 2, O 2, Ar, CO 2, and H 2 O) and initial trapped mass (m 0 ), including trapped residual mass Post-combustion composition determined assuming complete combustion of delivered fuel mass. Minor species resulting from dissociation during combustion not considered 25 CEFRC1-2, 2014

26 Part 2: Turbochargers, Engine Performance Metrics Herold, 2011 First law energy balance: de=dq - Pdv Combustion: Wall heat transfer: Combustion model - Wiebe function Heat transfer model - Woschni 26 CEFRC1-2, 2014

27 BTE [%] Part 2: Turbochargers, Engine Performance Metrics Chen-Flynn, 1965 Engine brake thermal efficiency BTE BTE*LHV=IMEPg-PMEP-FMEP DOE goal BTE=55% Friction model Chen-Flynn model ( SAE ). FMEP = C + (PF*P max ) + (MPSF*Speed mp ) where: + (MPSSF*Speed mp2 ) C = constant part of FMEP (0.25 bar) PF = Peak Cylinder Pressure Factor (0.005) P max = Maximum Cylinder Pressure MPSF = Mean Piston Speed Factor (0.1) MPSSF = Mean Piston Speed Squared Factor (0) Speed mp = Mean Piston Speed bar PCP Limit PMEP FMEP BTE GIE{1 } IMEPg UW Dyno limit PMEP = 0.4 bar FMEP = 1 bar UW RCCI GIE = 55% 30 SCOTE GIE = 60% results (Exp/Sim) GIE = 65% Load -- Gross IMEP [bar] 27 CEFRC1-2, 2014

28 Part 2: Turbochargers, Engine Performance Metrics Lavoie, D modeling for engine performance analysis 28 CEFRC1-2, 2014

29 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Mid load 29 CEFRC1-2, 2014

30 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Woshni, 1967 Turbocharger equation Burn duration Heat transfer Friction m~0.8, Re increases with Bore and (boost) 30 CEFRC1-2, 2014

31 Cumulative heat release Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Effect of combustion phasing on efficiency Constant volume combustion Burn 100% 90% 50% CA50 10% Crank angle Without HT: Best efficiency CA50~TDC With HT: best efficiency with CA50~10 deg tradeoff between heat loss/late expansion 31 CEFRC1-2, 2014

32 Energy budget Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 F0 air standard efficiency 63% Adiabatic Decreasing Incomplete combustion 32 CEFRC1-2, 2014

33 Burned gas temperature Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Effect of dilution Fuel-to-charge equivalence ratio, f f ranges from 0.2 to 1 with air, EGR ranges from 0 to 80% with f=1 33 CEFRC1-2, 2014

34 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Effect of boost on efficiency Reduced heat transfer loss Reduced friction losses 34 CEFRC1-2, 2014

35 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Potential brake efficiencies of naturally aspirated engines Increased pumping losses 35 CEFRC1-2, 2014

36 Summary Part 2: Turbochargers, Engine Performance Metrics Turbocharging can increase engine efficiency by using available energy in exhaust and by reducing pumping work Air standard ideal cycle analysis provides a bound on engine efficiency estimates. 0-D engine system models provide estimates of engine system efficiencies, if combustion details (e.g., timing and duration) and heat transfer losses are assumed The goal of multi-dimensional models (to be discussed next) is to predict the combustion details 36 CEFRC1-2, 2014

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