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 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 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 Part 2: Turbochargers, Engine Performance Metrics Turbocharging 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 Part 2: Turbochargers, Engine Performance Metrics
Turbocharging
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 Part 2: Turbochargers, Engine Performance Metrics Regulated two-stage turbocharger Duplicated Configuration per Cylinder Bank
LP stage Turbo-Charger with Bypass
Compressor HP stage Turbo Bypass charger
Charge Air Regulating valve Cooler
EGR Cooler
EGR Valve GT-Power R2S Turbo Circuit
HP TURBINE EGR Valve Compressor Bypass
EGR Cooler
Charge Air Regulating valve Cooler Compressor HP stage Turbo Bypass charger
LP stage Turbo-Charger with Bypass Regulating Valve
LP Stage Bypass LP TURBINE
6 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Intercooler for IVC temperature control Q P VIVC Isentropic PV IVC Reduced Peak Temp (NOx) Improved phasing
( 1) ln P T VIVC TVIVC ln T
Pressure Tign /time of Compressor ignition
Boost Q
TDC IVC TDC IVC ln V ln V Boost explains 20% of the improved fuel efficiency of diesel vs. SI 7 CEFRC1-2, 2014 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 Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990
Compressible flow – A review Area-velocity relations Gibbs Tds dh dp / for M<1 for M>1
Energy dh VdV
Euler dP VdV
d dA dV AV Const 0 AV 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 dA dV (M 2 1) AV dA(1 M 2 ) dP AV 2
Traffic flow behaves like a supersonic flow!
9 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990
Model passages as compressible flow in converging-diverging nozzles
PV m AV A RT RT c A* Minimum area point
1/ 2 P0 P0 AM( P / P 0 ) /( T / T 0 ) RT0
With M=1: Fliegner’s formula Choked flow, M=1
1 1 2 2( 1) * mM 10 () P A A*/A 1 RT0 Subsonic Supersonic
Area Mach number relations 2 solutions for 1 same area 2( 1) A 1 2 ( 1) 2 * (1M ) AM 12
1 1/ 2 11 0 APP21 1 1 reservoir 0.528 throat P/P exit 0 * ( ) 1 ( ) 0 APP00 12 0 1 M ∞
10 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990
Isentropic nozzle flows
T0 1 2 P0 1 2 1 1 M1 (1M1 ) Ex. Flow past throttle plate T1 2 P1 2
P0 P1 y
0 P 1 P=Pb 0 Choked flow for P2 < 53.5 kPa = 40.1cmHg reservoir ambient WOT
Choked
m 1 Pb P/P0 y 0.528 40.1 76 M=1 0 Manifold pressure, P1 cmHg x
11 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Anderson, 1990
Application to turbomachinery
Fliegner’s Formula:
1 2 2( 1) * Variable Geometry Compressor/ mM 10 () P A 1 RT0 turbine performance map
Increased speed
Choked flow “Corrected mass m Tref / T0
flow rate” PP0 / ref A measure of effective flow area Reduced flow passage area
1.0 1/0.528=1.89
P0 /P Total/static pressure ratio 12 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988 Compressor
(Toutisen Tin ) c (Tout Tin )
P0 3 T P3 = Pout Heywood, Fig. 6-43 Air at stagnation state 0,in accelerates to P2 inlet pressure, P1, and velocity V1.
Compression in impeller passages
increases pressure to P2, and velocity V2.
P0 = P 0,in Diffuser between states 2 and out, recovers air kinetic energy at exit of impeller 2 P1 V1 /2 c P producing pressure rise to, Pout and low velocity Vout
Wc m a h out h in S a 1 a Note: use exit static pressure and inlet total ma c P T in p W a out 1 pressure, because kinetic energy of gas c cp0, in leaving compressor is usually not recovered 13 CEFRC1-2, 2014 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 Speed/pressure limit line – reduced mass flow due to periodic flow reversal/reattachment in Non-dimensionalize blade passage boundary layers. tip speed (~ND) by speed Unstable flow can lead of sound to damage At high air flow rate, operation is limited by choking at the minimum Pressure ratio evaluated area point within compressor using total-to-static pressures since exit flow Supersonic flow kinetic energy is not recovered Shock wave Heywood, Fig. 6-46 14 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Serrano, 2007
Compressor maps
3.0 GM 1.9L diesel engine Pressure Ratio (t/t) 2.8 190000 35000 40000 50000 70000
2.6 90000 110000 130000 150000
170000 180000 190000 2.4
2.2 Efficiency 0.8 (T/T) 2.0 180000
0.7 170000 1.8
150000 0.6 1.6
130000 Corrected Air Flow (kg/s) 1.4 0.5 0.00 0.02 0.04 0.06 0.08 0.10 0.12 0.14 0.16 0.18 110000 1.2 90000
50000 70000 Corrected Air Flow (kg/s) 1.0 35000 40000 0.00 0.02 0.04 0.06 0.08 0.10 0.12 0.14 0.16 0.18
15 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Reitz, 2007 Automotive turbines Naturally aspirated:
Pintake=Pexhst=Patm (5-7-8-9-1) Boosted operation: Negative pumping work: Wt m g() h in h0, out P7 2 Expansion Blowdown 5 Compression Available work (area 5-6-7) 9 1 Pintake 6’’ Turbine P 6 Compressor exhst 8 7 6’ Pamb TDC BDC V P-V diagram showing available exhaust energy - turbocharging, turbocompounding, bottoming cycles and thermoelectric generators further utilize this available energy 16 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Turbochargers Radial flow – automotive; axial flow – locomotive, marine P0 = P0,in T P1 2 V1 /2 c P T3 P2 T0 mcorrected mg p3 p0 N out Ncorrected T3 P0 3 T0 P3 = Pout (Tout Tin ) t (Toutisen Tin ) S 17 CEFRC1-2, 2014 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 Part 2: Turbochargers, Engine Performance Metrics Engine breathing lines Engine Breathing Lines 1.4L Diesel, Air-to-Air AfterCooled, Turbocharged 3.8 3.6 Torque Peak (1700rpm) Trq Peak Operating Pnt 3.4 Rated (2300rpm) 3.2 Rated Operating Pnt 3 2.8 2.6 2.4 2.2 2 Compressor Pressure Ratio CompressorPressure 1.8 1.6 Parameter Torque Peak Rated Units 1.4 Horsepower 48 69 hp BSFC 0.377 0.401 lb/hp-hr 1.2 A/F 23.8 24.5 none 1 0.000 1.000 2.000 3.000 4.000 5.000 6.000 7.000 8.000 9.000 10.000 11.000 12.000 13.000 14.000 Intake Mass Flow Rate (lb/min) 19 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988 . . Wt = Wc a a1 g 1 g p2 Cpg T3 m fuel p4 1 1 t c mech 1 p Cp T p 1 a 1 m 3 air 20 CEFRC1-2, 2014 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 Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988 Ideal engine efficiency – Otto cycle Otto Efficiency = net work / energy supplied [(T 3 T 4) (T 2 T1)]/(T 3 T 2) T 3 1(T 4 T1)/(T 3 T 2) 2 Wexpansion However, Wcompression 1 1 1 T2/(/)(/)/ T 1 V 1 V 2 rc V 4 V 3 T 3 T 4 4 1 0.8 =1.4 1.3 0.6 s 1.25 1 0.4 11/ rc 0.2 8 16 24 0 rc 22 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics ηideal Function of only two variables, compression ratio (rc) and ratio of specific heats (γ) Increasing rc 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 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 0.65. 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 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 (θSOI) and injection duration (Δθinj). Pressure and mass integrated over the closed portion of cycle with specified initial conditions at IVC of pressure (p0), temperature (T0), and composition (xn,0 for all species considered - N2, O2, Ar, CO2, and H2O) and initial trapped mass (m0), 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 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 Part 2: Turbochargers, Engine Performance Metrics Chen-Flynn, 1965 70 Engine brake thermal efficiency BTE PMEP FMEP BTE GIE{1 } 150 bar PCP Limit IMEPg BTE*LHV=IMEPg-PMEP-FMEP 60 DOE goal BTE=55% 55 50 UW Dyno limit 45 Friction model PMEP = 0.4 bar BTE[%] 40 FMEP = 1 bar Chen-Flynn model ( SAE 650733). UW RCCI GIE = 55% 30 SCOTE FMEP = C + (PF*Pmax) + (MPSF*Speedmp) GIE = 60% results (Exp/Sim) GIE = 65% 2 + (MPSSF*Speedmp ) 20 0 5 10 15 20 25 30 where: C = constant part of FMEP (0.25 bar) Load -- Gross IMEP [bar] PF = Peak Cylinder Pressure Factor (0.005) Pmax = Maximum Cylinder Pressure MPSF = Mean Piston Speed Factor (0.1) MPSSF = Mean Piston Speed Squared Factor (0) Speedmp = Mean Piston Speed 27 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 1-D modeling for engine performance analysis 28 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Mid load 29 CEFRC1-2, 2014 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 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Effect of combustion phasing on efficiency Constant volume combustion 10-90 Burn 100% 90% 50% CA50 10% Cumulative release heat Cumulative 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 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Energy budget F0 air standard efficiency 63% Adiabatic Decreasing Incomplete combustion 32 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Effect of dilution Fuel-to-charge equivalence ratio, f’ Burned temperature gas Burned f ranges from 0.2 to 1 with air, EGR ranges from 0 to 80% with f=1 33 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Effect of boost on efficiency Reduced heat transfer loss Reduced friction losses 34 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Lavoie, 2012 Potential brake efficiencies of naturally aspirated engines Increased pumping losses 35 CEFRC1-2, 2014 Part 2: Turbochargers, Engine Performance Metrics Summary 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