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Turbochargers, Engine Performance Metrics

Turbochargers, Engine Performance Metrics 14
Part 2: Turbochargers, Engine Performance Metrics Reciprocating Internal Combustion Engines Prof. Rolf D. Reitz Engine Research Center University of WisconsinMadison 2014 PrincetonCEFRC 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 CEFRC12 , 2014 Part 2: Turbochargers, Engine Performance Metrics Short course outine: Engine fundamentals and performance metrics, computer modeling supported by indepth understanding of fundamental engine processes and detailed experiments in engine design optimization. Day 1 (Engine fundamentals) Part 1: IC Engine Review, 0, 1 and 3D 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, Aftertreatment and Controls Part 10: Vehicle Applications, Future of IC Engines 2 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics Turbocharging Pulsedriven 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 CEFRC12, 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 twostrokes for effective scavenging: intake P exhaust P crankcase used as a pump Some engines combine enginedriven and mechanical (e.g., in twostage configuration). Intercooler after compressor controls combustion air temperature. 4 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics Turbocharging Energy in exhaust is used to drive turbine which drives compressor Wastegate used to bypass turbine Charge air cooling after compressor further increases air density more air for combustion 5 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics Regulated twostage turbocharger Duplicated Configuration per Cylinder Bank LP stage TurboCharger with Bypass HP stage Turbo Compressor charger Bypass Charge Air Regulating valve Cooler EGR Cooler EGR Valve GTPower R2S Turbo Circuit HP TURBINE Compressor Bypass EGR Valve EGR Cooler Charge Air Regulating valve Cooler Compressor HP stage Turbo Bypass charger LP stage TurboCharger with Bypass Regulating Valve LP Stage Bypass LP TURBINE 6 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics Intercooler for IVC temperature control Q  P V  IVC Isentropic   PV  IVC Reduced Peak Temp (NOx) Improved phasing ( 1) T V  ln P IVC   TV  IVC ln T Pressure T ign /time of Compressor ignition Boost Q IVC TDC IVC TDC ln V ln V Boost explains 20 of the improved fuel efficiency of diesel vs. SI 7 CEFRC12, 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 exhaustdriven 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 CEFRC12, 2014 Anderson, 1990 Part 2: Turbochargers, Engine Performance Metrics Compressible flow – A review Areavelocity relations Tds dh dp /  Gibbs  for M1 for M1 Energy dhVdV  dPVdV Euler  d  dA dV  AV  Const    0 Subsonic nozzle Subsonic diffuser Supersonic diffuser Supersonic nozzle  AV dA0 dA 0 dA 0 dA 0 from AV  dV0 dV 0 dV 0 dV 0 from Euler  dP0 dP 0 dP 0 dP 0 kinetic energy pressure recovery kinetic energy dA dV 2  (M 1) AV 2 dA (1 M )  dP 2 AV  Traffic flow behaves like a supersonic flow 9 CEFRC12, 2014 Anderson, 1990 Part 2: Turbochargers, Engine Performance Metrics Model passages as compressible flow in convergingdiverging nozzles PV m AV A RT Minimum area point A RT c P  1/ 2 0  P AM (P / P ) /(T /T ) 0 0 0 RT 0 With M=1: Fliegner’s formula Choked flow, M=1 1  1 2  2( 1) m () P A A/A M 10  1 RT 0 Subsonic Supersonic 2 solutions for Area Mach number relations  1 same area 2( 1) A 1 2 ( 1) 2  (1 M )  AM  12  1/ 2  1 11   0   1 A P21 P     0.528 0 1  reservoir P/P throat  ( ) 1 ( ) exit  0   A P  12 P   00  0 1 M ∞ 10 CEFRC12, 2014 Anderson, 1990 Part 2: Turbochargers, Engine Performance Metrics Isentropic nozzle flows  T  1 P  1 2 2  1 0 0  1 M  (1 M ) 1 1 Ex. Flow past throttle plate P 2 T 2 1 1 P P 0 1 y 1 0 P=P P b 0 Choked flow for P 53.5 kPa = 40.1cmHg 2 ambient reservoir WOT Choked m 1 P b P/P 0 y 0.528 40.1 76 M=1 Manifold pressure, P cmHg 1 0 x 11 CEFRC12, 2014 Anderson, 1990 Part 2: Turbochargers, Engine Performance Metrics Application to turbomachinery Fliegner’s Formula:  1 2  Variable Geometry Compressor/ 2( 1) m () P A M 10 turbine performance map  1 RT 0 Increased speed Choked flow m T /T “Corrected mass ref 0 flow rate” PP / 0 ref A measure of effective flow Reduced flow passage area area 1.0 1/0.528=1.89 P /P 0 Total/static pressure ratio 12 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988 Compressor (T T ) outisen in   c (T T ) out in P 03 T P = P 3 out Heywood, Fig. 643 Air at stagnation state 0,in accelerates to P 2 inlet pressure, P , and velocity V . 1 1 Compression in impeller passages increases pressure to P , and velocity V . 2 2 P = P 0 0,in Diffuser between states 2 and out, recovers air kinetic energy at exit of impeller P 2 1 producing pressure rise to, P and V /2 c out 1 P low velocity V out W m h h   c a out in  1 a S   a m c T  p a P in Note: use exit static pressure and inlet total a out W1  c   pressure, because kinetic energy of gas  p c 0,in   leaving compressor is usually not recovered  13 CEFRC12, 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 Nondimensionalize 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 totaltostatic pressures since exit flow Supersonic flow kinetic energy is not recovered Shock wave Heywood, Fig. 646 14 CEFRC12, 2014 Serrano, 2007 Part 2: Turbochargers, Engine Performance Metrics Compressor maps 3.0 Pressure GM 1.9L diesel engine Ratio (t/t) 2.8 190000 35000 40000 50000 70000 2.6 90000 110000 130000 150000 2.4 170000 180000 190000 2.2 Efficiency 0.8 (T/T) 2.0 180000 170000 0.7 1.8 150000 0.6 1.6 130000 Corrected Air Flow (kg/s) 1.4 0.5 110000 0.00 0.02 0.04 0.06 0.08 0.10 0.12 0.14 0.16 0.18 1.2 90000 70000 Corrected Air Flow (kg/s) 50000 35000 40000 1.0 0.00 0.02 0.04 0.06 0.08 0.10 0.12 0.14 0.16 0.18 15 CEFRC12, 2014 Reitz, 2007 Part 2: Turbochargers, Engine Performance Metrics Automotive turbines Naturally aspirated: P =P =P (57891) intake exhst atm Boosted operation: Negative pumping work: W m() h h t g in 0,out P P – but hurts scavenging 7 1  1 g  P  g  P  0,out W m c T  1  t g P in t  3 4 P in    Expansion 2 Blowdown 5 Available work Compression (area 567) 1 9 6’’ P Turbine intake 6 P exhst Compressor 8 7 6’ P amb BDC TDC V PV diagram showing available exhaust energy turbocharging, turbocompounding, bottoming cycles and thermoelectric generators further utilize this available energy 16 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics Turbochargers Radial flow – automotive; axial flow – locomotive, marine P 0 = P 0,in T P 1 2 V /2 c 1 P T 3   T 0 P 2 m  m corrected g p 3 p 0 out N N  corrected T 3 P T 0 3 0 P = P 3 out (T T ) out in   t (T T ) outisen in S 17 CEFRC12, 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 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics Engine breathing lines Engine Breathing Lines 1.4L Diesel, AirtoAir 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 1.8 1.6 Parameter Torque Peak Rated Units 1.4 Horsepower 48 69 hp BSFC 0.377 0.401 lb/hphr 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 CEFRC12, 2014 Compressor Pressure RatioPart 2: Turbochargers, Engine Performance Metrics Heywood, 1988 . . W = W t c  a  1 a  1  g        Cp T m  g      p   p   g 3 fuel 2 4      1 1     1   t c mech          p Cp T p    1  a 1  3    m   air       20 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics Heywood, 1988 Ideal engine efficiency – Otto cycle Maximum possible closedcycle 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 CEFRC12, 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 W expansion However, W compression   1   1   1 T 2 /T1  (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  11/ rc 0.2 8 24 16 0 r c 22 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics η Function of only two variables, compression ratio (r ) ideal c and ratio of specific heats (γ) Increasing r increases operating volume for compression and expansion c 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 closedcycle 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 CEFRC12, 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 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics Herold, 2011 Zerodimensional closedcycle analysis: Combustion represented as energy addition to closed system Fuel injection mass addition from userspecified start of injection crank angle (θ ) and injection duration (Δθ ). SOI inj Pressure and mass integrated over the closed portion of cycle with specified initial conditions at IVC of pressure (p ), temperature (T ), and composition 0 0 (x for all species considered N , O , Ar, CO , and H O) and initial trapped n,0 2 2 2 2 mass (m ), including trapped residual mass 0 Postcombustion composition determined assuming complete combustion of delivered fuel mass. Minor species resulting from dissociation during combustion not considered 25 CEFRC12, 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 CEFRC12, 2014 Part 2: Turbochargers, Engine Performance Metrics ChenFlynn, 1965 70 Engine brake thermal efficiency BTE PMEP  FMEP BTE GIE1 IMEPg 150 bar PCP Limit BTELHV=IMEPgPMEPFMEP 60 55 DOE goal BTE=55 50 UW Dyno limit 45 Friction model PMEP = 0.4 bar 40 FMEP = 1 bar ChenFlynn model ( SAE 650733). UW RCCI GIE = 55 30 SCOTE FMEP = C + (PFP ) + (MPSFSpeed ) max mp GIE = 60 results GIE = 65 (Exp/Sim) 2 + (MPSSFSpeed ) 20 mp 0 5 10 15 20 25 30 Load Gross IMEP bar where: C = constant part of FMEP (0.25 bar) PF = Peak Cylinder Pressure Factor (0.005) P = Maximum Cylinder Pressure max MPSF = Mean Piston Speed Factor (0.1) MPSSF = Mean Piston Speed Squared Factor (0) Speed = Mean Piston Speed mp 27 CEFRC12, 2014 BTE Lavoie, 2012 Part 2: Turbochargers, Engine Performance Metrics 1D modeling for engine performance analysis 28 CEFRC12, 2014 Lavoie, 2012 Part 2: Turbochargers, Engine Performance Metrics Mid load 29 CEFRC12, 2014 Lavoie, 2012 Part 2: Turbochargers, Engine Performance Metrics Woshni, 1967 Turbocharger equation Burn duration Heat transfer m0.8, Re increases with Bore and  (boost) Friction 30 CEFRC12, 2014 Lavoie, 2012 Part 2: Turbochargers, Engine Performance Metrics Effect of combustion phasing on efficiency Constant volume combustion 1090 Burn 100 90 50 CA50 10 Crank angle Without HT: Best efficiency CA50TDC With HT: best efficiency with CA5010 deg – tradeoff between heat loss/late expansion 31 CEFRC12, 2014 Cumulative heat release Lavoie, 2012 Part 2: Turbochargers, Engine Performance Metrics Energy budget F0  air standard efficiency 63 Adiabatic Decreasing  Incomplete combustion 32 CEFRC12, 2014 Lavoie, 2012 Part 2: Turbochargers, Engine Performance Metrics Effect of dilution Fueltocharge equivalence ratio, f’ f ranges from 0.2 to 1 with air, EGR ranges from 0 to 80 with f=1 33 CEFRC12, 2014 Burned gas temperature Lavoie, 2012 Part 2: Turbochargers, Engine Performance Metrics Effect of boost on efficiency Reduced heat transfer loss Reduced friction losses 34 CEFRC12, 2014 Lavoie, 2012 Part 2: Turbochargers, Engine Performance Metrics Potential brake efficiencies of naturally aspirated engines Increased pumping losses 35 CEFRC12, 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. 0D 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 multidimensional models (to be discussed next) is to predict the combustion details 36 CEFRC12, 2014