F7020T Internal Combustion Engines PDF

LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE F7020T Internal Combustion Engines O1 – Introduction  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Internal combustion engines (ICEs) Combustion of a fuel occur...

LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE F7020T Internal Combustion Engines O1 – Introduction  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Internal combustion engines (ICEs) Combustion of a fuel occurs with an oxidizer (usually air) in a combustion chamber, and the high-temperature high-pressure combustion products apply a direct force on some component in order to convert fluid energy into mechanical energy This force is typically applied to: - pistons (reciprocating or rotary ICEs) - blades (gas turbines, jet engines) - nozzles (rocket engines) On the contrary, in external combustion engines (e.g., Stirling engines) the working fluid is not contaminated by combustion products. In this course only reciprocating ICEs will be covered (commonly the term “ICE” refers to this specific kind of engines).  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE History Before the industrial revolution the only sources of mechanical power were water- and windmills XVIII century - Newcomen’s and Watt’s atmospheric steam engines 1790s compressionless engines (Leonardo da Vinci already described one) 1800 Trevithick and Evans introduce high-pressure steam engines 1824 Carnot’s thermodynamic theory on idealized heat engines: compression is necessary to differentiate high and low working temperatures 1854-57 Barsanti/Matteucci describe the principles of atmospheric engines 1860s Lenoir and Otto/Langen develop commercial atmospheric engines 1862 Beau de Rochas describe the principles of the 4-stroke engine 1876 Otto develops the first modern 4-stroke engine 1880s Clerk, Robson and Benz develop 2-stroke engines 1892 Diesel patents his compression-ignition engine (built 5 years later at MAN) 1957 Wankel successfully tests his rotary engine (patented in 1929) at NSU  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Classification Application: automobile, truck, locomotive, light aircraft, marine, portable system power generation Basic design engines: arrangement of cylinders (in-line, V, radial, opposed) Working cycle: 2-stroke, 4-stroke, (6-stroke) Valve or port design and location Fuel: gasoline (or petrol), diesel fuel (or fuel oil), natural gas, LPG, alcohols, biofuels, hydrogen, dual fuel Method of mixture preparation: carburation, indirect or direct injection Method of ignition: spark ignition or compression ignition Combustion chamber design Method of load control: both air and fuel flows (throttling) or fuel flow only Method of cooling: water cooled, air cooled, natural convection and radiation  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Classification  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE 4-stroke engine cycle 1st stroke 2nd stroke 3rd stroke 4th stroke intake compression power or exhaust expansion  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE 2-stroke engine cycle 1st stroke 2nd stroke expansion and exhaust intake and compression (compression in crankcase) (crankcase filling)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Engine main mechanical components Engine block: contains cylinders and cooling passages, cylinder sleeves may be removable. Crankcase (bottom part) is usually the tank for lubrication oil Crankshaft: supported by main bearings  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Engine main mechanical components Connecting rod Piston (piston pin, rings, skirt)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Engine main mechanical components Engine head: contains intake and exhaust valve ports, guides and seats, spark plugs, fuel injectors Valve train: camshafts, poppet valves and return springs LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE F7020T Internal Combustion Engines O2 – Reference and real operating cycles  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Thermodynamic cycles Sequences of transformations in which the operating fluid exchanges heat and work with the external environment, the initial and the final state coincide A cycle is direct if the net effect is transferring work to the environment, inverse if work is spent to transfer thermal energy from low to high temperature First and second law of thermodynamics establish the fundamental features of a direct cycle: T1 1st : W = Q1 - |Q2| Q1 2nd : W < Q1 W  = W / Q1= (Q1 - |Q2|) / Q1 Q2 T2 thermal efficiency  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Why thermodynamic cycles ? In internal combustion engines the operating fluid undergoes a chemical transformation that replaces the thermodynamic transformation in the operating fluid receives heat from the hot temperature source -> “fresh” operating fluid must substitute the “exhausted” one at the end of every cycle -> Internal combustion engines do not operate according to thermodynamic cycles! Reference basic thermodynamic cycles are associated with engine operation only to obtain some approximated estimation about the performance and the influence of the main operating parameters (e.g., pressure and temperature ratios) on performance  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Why thermodynamic cycles ? ICE operation can be described with different approximation levels by: - perfect gas cycles operating fluid is air, considered as a perfect gas (cp, cv,  = const.) heat is exchanged instantaneously and with external sources/sinks – no heat losses all transformation are reversible - air cycles operating fluid is air, considered as an ideal gas (cp, cv,  = f(T)) -> lower cycle pressures and temperatures -> lower thermal efficiency - fuel-air cycles operating fluid is the actual mixture of chemical species inside the cylinder (cp, cv,  = f(T, composition)) -> even lower cycle pressures and temperatures -> lower thermal efficiency The real operating cycle is determined experimentally, by measuring pressure and volume inside the cylinder. The main differences are due to: - heat losses (-> lower cycle pressures and temperatures) - fluid friction losses inside the operating fluid (-> less power obtained in the expansion) - the duration of the combustion process - valve opening and closure is not instantaneous  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Some well-known thermodynamics Thermodynamic transformations can be represented in the diagrams p-v and T-s pressure p temperature T 1 2 2 1 W12 Q12 specific volume v = V/m specific entropy s 2 2 dW = p dv → W12 =  p dv dQ = T ds → Q12 = T ds 1 1 dv > 0 -> expansion, fluid makes work ds > 0 -> fluid is heated dv < 0 -> compression, work done on the fluid ds < 0 -> fluid is cooled p T For a cyclic sequence of transformations the diagrams will show a closed area, W Q with W=Q v s  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Polytropic transformations All the reversible thermodynamic transformations used to build the reference cycles for ICEs are polytropic transformations as pvn = const. c=cv pressure p temperature T n=∞ c=cp n=0 c=∞ n=1 n= c=0 specific volume v = V/m specific entropy s - isobaric n=0 c=cp W=pv Q=cpT - isochoric n=∞ c=cv W=0 Q=cvT - adiabatic n= =cp/cv c=0 (dQ=0) W=cvT Q=0 ( isothermal n=1 c=∞ (dT=0) ) polytropic exponent specific heat (dQ/dT)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Maximum thermal efficiency Carnot established the maximum thermal efficiency for an ideal cycle operating with two capacities at different temperature (T1 and T2) Q1 pressure p 1’ temperature T 1 Q1 1 1’ T1 Q=W W 2’ Q2 T2 2’ 2 2 Q2 specific volume v = V/m specific entropy s The cycle with maximum thermal efficiency is made of 2 adiabatic and 2 isothermal transformations: (to be compared with the T2 T2' 1 1 id ,max = 1 − = 1 − = 1 −  −1 id ,max = 1 −  −1 thermal efficiency of the T1 T1  v2 '  r reference cycles for ICEs)    v1  r = v2’/v1 = compression ratio of the adiabatic transformation 1-2’  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Otto reference cycle (spark ignition engines) 0 -> 1 intake stroke 3 3 pressure p temperature T 1 -> 2 adiabatic compression |Win|=cv(T2-T1) Q=0 Qin 2 -> 3 isochoric combustion W=0 Qin=cv(T3-T2) 4 3 -> 4 adiabatic expansion 2 Q=W Wout=cv(T3-T4) Q=0 pa 1 4 -> 1 “isochoric exhaust” Qout W=0 |Qout|=cv(T4-T1) specific entropy s 1 -> 0 exhaust stroke W W = Q = Wout − Win = Qin − Qout = 2 (BDC) = cv (T3 − T4 ) − cv (T2 − T1 ) = cv (T3 − T2 ) − cv (T4 − T1 ) (TDC) 4 pa 0 1 Qin − Qout cv (T3 − T2 ) − cv (T4 − T1 ) T −T id = = =1− 4 1 = Vc/m Vd/m specific volume v Qin cv (T3 − T2 ) T3 − T2  −1 V +V T (T T − 1) T v  1 v =1− 1 4 1 = 1 − 1 = 1 −  2  =1− r = c d = 2 (m = const.) Vc v1 T2 (T3 T2 − 1) T2  v1  r  −1 The thermal efficiency of the ideal Otto cycle is equal to the efficiency of a Carnot cycle between T1 and T2 (instead of the maximum temperature of the cycle, that is T3)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Diesel ref. cycle (compression ignition engines) combustion transformation is less steeper than in Otto 0 -> 1 intake stroke cycle (cv 2 adiabatic compression r is higher Qin 3 |Win|=cv(T2-T1) Q=0 than in 2 -> 3 isobaric combustion Otto cycle 2 W’out=p2(v3-v2) Qin=cp(T3-T2) Q=W 4 3 -> 4 adiabatic expansion 2 3 W’’out=cv(T3-T4) Q=0 pa 1 4 -> 1 “isochoric exhaust” Qout W=0 |Qout|=cv(T4-T1) specific entropy s 1 -> 0 exhaust stroke W Qin − Qout c p (T3 − T2 ) − cv (T4 − T1 ) 1 T4 − T1 (BDC) id = = = 1− = (TDC) Qin c p (T3 − T2 )  T3 − T2 4 0 1  1   −1  v3 T3 pa 1 = 1 −  −1    with = = Vc/m Vd/m specific volume v r    − 1  v2 T2 The thermal efficiency of the ideal Diesel cycle is lower than that of the ideal Otto cycle as  > 1  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Dual reference cycle 0 -> 1 intake stroke 1 -> 2 adiabatic compression Q’’in 4 |Win|=cv(T2-T1) Q=0 pressure p temperature T 3 2 -> 3 isochoric combustion 3 4 Q’in W=0 Q’in=cv(T3-T2) 3 -> 4 isobaric combustion 2 W’out=p3(v4-v3) Q’’in=cp(T4-T3) Q=W 5 4 -> 5 adiabatic expansion 2 W’’out=cv(T4-T5) Q=0 pa 1 Qout 5 -> 1 “isochoric exhaust” W=0 |Qout|=cv(T5-T1) specific entropy s 1 -> 0 exhaust stroke W Q 'in +Q ''in − Qout cv (T3 − T2 ) + c p (T4 − T3 ) − cv (T5 − T1 ) (TDC) (BDC) id = = = Q 'in +Q ''in cv (T3 − T2 ) + c p (T4 − T3 ) 5 0 1 1   −1 p3 v4 T4 pa = 1 −  −1 with = = = Vc/m Vd/m specific volume v r ( −1) +  (  − 1) p2 v3 T3 Real combustion is too fast to be considered a constant pressure process, and too slow to be considered an instantaneous constant volume process -> dual cycle is used to approximate both SI and CI engine cycles With  =1 -> Diesel reference cycle with  =1 -> Otto reference cycle  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Comparison among reference cycles ideal thermal efficiency id SI same r engines same Qin CI engines Otto 12341 – Diesel 123’4’1 – dual 122’’3’’4’’1 compression ratio r Ideal thermal efficiency always increases with compression ratio r id,Otto > id,dual > id,Diesel , but in SI engines r is limited to avoid knock -> CI engines have higher thermal efficiency This is also shown in a comparison with same r and same Qin. The cycle which rejects more heat is the one with the lower thermal efficiency, and indeed Area(6145) (Otto) < Area(614’’5’’) (dual) < Area(614’5’) (Diesel)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Atkinson-Miller reference cycle 3 pressure p Since p4>pa, more work can be obtained if the adiabatic expansion is prolonged to pa -> Atkinson-Miller cycle This however requires a larger displacement volume (v4’>v4=v1) and it is apparent that rexpansion>rcompression This cycle can be operated in the same cylinder only if valves are left open between 4’ and 1 during the compression stroke (the actual compression ratio is less than the geometrical one as some of the aspirated mixture flows back in the intake duct) W 2 4 pa 0 1 4’ specific volume v (TDC) (BDC comp.) (BDC exp.) - Thermal efficiency increase is significant only if r is low - The work obtained for the same displacement volume is lower than in the other reference cycles, because the actual r is lower than the geometrical one - The gas exchange process (the discharge of the combustion gases in particular) is easier if cylinder pressure is higher than ambient pressure when the exhaust valves are opened (i.e. if p4>pa)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Real operating cycle The real operating cycle is determined experimentally by measuring the pressure inside the cylinder as the volume varies due to piston motion The pressure diagram is also called “indicated” cycle because of the name the “indicator” instrument that was used to draw it. specific volume v Rq : mass isn't consistent during real cycles (valves opening) => we use V=A*x instead of v=V/m  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Real SI and CI engine operating cycles The shape of real SI and CI engine operating pressure p cycles does not differ much: - a higher maximum pressure is reached by the CI engine due to the higher compression ratio (20-24 vs 11-12) - a small step near the end of the compression stroke can be seen in the CI engine cycle if ignition delay is long CI engine SI engine Note that real operating cycles are made of two areas: atm. press. - a main area (clockwise path) providing work - a smaller area (counterclockwise path) absorbing work imep’ imep which corresponds to the pumping work required during the exhaust and the intake strokes for the gas exchange process CI engine volume V SI engine Vd The overall net work is given by the difference of the two areas (indicated work Wi) Comparison between SI and CI engine Although this is not a thermodynamic cycle, the area in the operating cycles. The displacement volume diagram still represents work since of the two cycles has been superimposed, dW = F dx = p Ap dx = p dV since the volume of the combustion chambers is different (r CI > rSI )  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Real SI and CI engine operating cycles the indicated efficiency of the real operating pressure p cycle is obtained dividing the indicated work by the heat input: Wi i = = idic CI engine Q SI engine The ratio between the indicated efficiency and atm. press. the ideal reference cycle efficiency is an index of the differences between the indicated and imep’ imep the reference cycle, and it is called “indicated cycle efficiency” CI engine volume V Wi i SI engine Vd ic = = Wid id Comparison between SI and CI engine operating cycles. The displacement volume of the two cycles has been superimposed, But which are these differences? since the volume of the combustion chambers is different (r CI > rSI )  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Reference and real SI engine cycles A reference and a real SI engine cycle are reference cycle pressure p compared for the same compression ratio indicated cycle and heat input. The differences are marked with the letters A,B,C and D A) real properties of air/fuel mixture and combustion gases: exhaust valve - specific heats and their ratio increase with opens temperature -> maximum pressure and ignition temperature are lower in the real cycle - compression and expansion are not adiabatic due to heat losses (polytropic exhaust transformations) -> more work required for pa compression, less work obtained from intake expansion volume V (TDC) (BDC) - dissociation of combustion products at high temperature reduces maximum pressure  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Reference and real SI engine cycles B) combustion is not instantaneous: spark reference cycle pressure p ignition must occur before TDC is reached indicated cycle to approximate a constant volume process. The shape of the real cycle is rounded and some work is lost. The optimal spark advance is determined experimentally (compression work is abnormally raised if ignition occurs too early, high pressures cannot be reached if ignition occurs too late). exhaust valve Combustion can also be incomplete opens (combustion efficiency) ignition lost work exhaust lost work pa intake (TDC) volume V (BDC) normal ignition early ignition TDC BDC TDC BDC late ignition normal ignition  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Reference and real SI engine cycles C) exhaust valve timing: the valve is not reference cycle pressure p opened instantaneously at BDC. The valve indicated cycle must be opened in advance to reduce the pressure of the combustion gases before the exhaust stroke begins, otherwise the work spent during this stroke is excessive (the same if the resistance of the exhaust exhaust valve system to the flow is high). If the valve opens opens too early, some expansion work is ignition lost -> a compromise must be found. p p additional work lost expansion spent in exhaust work exhaust stroke pa intake (TDC) volume V (BDC) BDC V BDC V exhaust valve opens too late exhaust valve opens too early  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Reference and real SI engine cycles D) pumping work area: cylinder pressure is reference cycle pressure p higher than atmospheric pressure during indicated cycle the exhaust stroke, and lower during the intake stroke -> negative work area. This area increases when: - intake system resistance is high - intake valve opens too late - throttle valve is partially closed exhaust exhaust valve pa intake opens ignition TDC high resistance BDC intake lost work Full load Partial load (throttle (throttle exhaust fully open) partially closed) pa intake ignition ignition (TDC) volume V (BDC) exhaust exhaust valve opens valve opens exhaust exhaust pa TDC intake BDC TDC intake BDC  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Reference and real CI engine cycles reference cycle pressure p Similar remarks can be done for the indicated cycle comparison between a reference and a real CI engine cycle for the same compression ratio and heat input. injection A) real properties of air/fuel mixture and exhaust valve combustion gases opens B) combustion is not a constant pressure process (isobaric combustion is approximated only in large low speed engines - higher combustion efficiency due to air excess) C) exhaust valve timing exhaust pa D) pumping work area small high intake (no throttling -> no increase speed CI engine at partial load) volume V (BDC) (TDC) large low speed CI engine TDC BDC  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Pressure-crank angle diagram (4-stroke engines) Cylinder pressure can also be plotted vs. crankshaft angle. This diagram highlights that cycle processes do not coincide with piston strokes. TDC BDC TDC BDC TDC strokes intake compression power exhaust processes intake comp. comb.+exp. exhaust 1 - at TDC p>patm because of the exhaust phase 7 - maximum pressure is after TDC because an 4 - intake valve should close when p=patm, or with effective combustion phase must be astride the a little delay to exploit the inertia of the gas TDC column entering the cylinder 8 - exhaust valve open and blowdown (spontaneous 5 - spark ignition (the dashed line 5-6’-… is cylinder exhaust) occurs, perhaps followed by a rapid pressure history without ignition) depression (10) due to gas column inertia  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Pressure-crank angle diagram (2-stroke engines) Cylinder pressure can also be plotted vs. crankshaft angle. This diagram highlights that cycle processes do not coincide with piston strokes. 5 - exhaust port opens first to reduce the pressure inside the cylinder, but then closes (2’’) after the intake port (2’) The actual compression ratio is lower than the 2’’,5 2’,1 geometrical one 2 The work spent to pressurize the fresh charge is not shown in the indicated diagram BDC TDC BDC exhaust port exhaust port opens ignition opens intake port opens intake port opens, scavenging begins pa exh. port closes intake in. port closes Vd comp. comb.+exp. exhaust LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE F7020T Internal Combustion Engines O3 – Fundamental quantities  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Geometric and kinematic parameters The parameters describing the crank-connecting rod mechanism in a reciprocating internal combustion engine are: - bore b - crank radius a (stroke s = 2a) - connecting rod length l b Piston displacement, velocity and acceleration would be sinusoidal functions only if l /a → ∞ s l a s  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Geometric and kinematic parameters - piston area Ap= b2/4 - displacement volume Vd= sAp= b2s/4 - number of cylinders nc - total displacement V= nc Vd= nc b2s/4 - clearance volume Vc (volume at TDC) Vc - compression ratio r = (Vd + Vc )/ Vc (volume at BDC / volume at TDC) Vd - rotational speed: N (rounds per unit of time) or  (crankshaft angular velocity); =2N - operating cycle frequency 2N/ns (ns is the number of strokes) - crank angle  = t =2Nt - mean piston speed: the average velocity during one crankshaft revolution (the length of two strokes divided  by the time required by one crankshaft revolution) U p = 2 Ns  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Mean piston speed Is one of the fundamental parameters to compare engine performance, as performance scales with piston velocity (rotational speed does not take into account the stroke, that is space). Mean piston speed can be correlated with: - fluid dynamic losses in the gas exchange process - turbulence intensity of organized charge motions in the cylinder - inertia forces loading mechanism joints - mechanical friction (piston-rings-cylinder) - heat losses through cylinder walls Mean piston speed is limited by fluid dynamic losses or by mechanical stresses due to inertia forces and is typically between 8 and 18 m/s Two engines that are geometrically similar and have the same mean piston speed also have the same velocity field (but field points are spatially scaled) 50 m/s b = s = 40 mm N = 8000 rpm 15 m/s 50 m/s b = s = 80 mm 15 m/s N = 4000 rpm U p = 10.67 m/s U p = 10.67 m/s  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Power and torque Are the main performance parameters of an internal combustion engine: Torque  measures the ability of the engine to do work Power P measures the rate at which that work is done P =  = 2 N When different engines are compared, power can be related to other parameters: - specific power (power per unit volume): P / V - weight/power ratio: (engine mass) / P - power per piston area: P / Ap  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Brake and indicated quantities p Pi Pb engine The power developed by the gases inside the load cylinder (indicated cycle) is not the power available at the crankshaft, part of it is wasted internally in the engine V Pf Three categories of power can be distinguished: - indicated power Pi - friction power Pf Pi − Pb = Pf - brake power Pb Indicated mean effective pressure (imep) Is an index of the work done by the gases inside the cylinder, and is independent from engine size It is equal to the constant pressure that should be applied to piston area during the expansion stroke only in order to get the same work of the indicated cycle imep 2N  2N  Pi = imep V  = (imep Ap ) s nc  Vd ns  ns   LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Mechanical efficiency m Is defined as the ratio between brake and indicated power, but is also the ratio between any other brake and indicated quantity Pb m = Pi N [rps] It takes into account: U p [m/s] - mechanical friction - pumping work in the gas exchange process (-> gross and net indicated quantities) - the driving of engine accessories gross p indicated Mechanical efficiency decreases: Pi work - as speed increases at constant load (friction power increases) - as load decreases at constant speed Pb (load) (friction power is fairly constant, so the Pf its quota of indicated power. At zero V load all the indicated power is spent pumping work in friction power)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Brake mean effective pressure Indicated mean effective pressure (which is a “real” parameter) can be considered as the sum of two “fictitious” mean effective pressures: imep = bmep + fmep bmep =  m imep Brake mean effective pressure is the other fundamental parameter to compare engine performance, because it is independent from engine size (as mean piston speed). Good engine designs should have nearly the same bmep, as it is an index of how effectively the available displacement volume has been exploited 2N Pb = bmep V ns SI engines (naturally aspirated): 11-12 bar CI engines (naturally aspirated): 9-10 bar SI engines (supercharged): 15-17 bar CI engines (supercharged): 18-20 bar  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Efficiency chain Starting from the heat Q which is released in the cylinder by combustion (Q = mf qc ) Wid ideal work of the reference ideal thermal efficiency id = Q cycle Wid Wi i indicated cycle efficiency ic = = Wid id Wi indicated work Wi indicated thermal efficiency i = = idic Q Wb (b ) mechanical efficiency  m = = Wi i Wb brake work Wb brake thermal efficiency (b ) = = i m = idic m Q gas di scarico irraggiamento exhaust 35% radiation 15% 10% gases 35% ENGINE THERMAL BALANCE all'albero brake power 30% raffreddamento cooling 25% 30% 20%  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Other important parameters Brake specific fuel consumption (bsfc) m f P P 1 bsfc =  = b = b = qc = fuel lower heating value [J/kg] P b Q m q bsfc q f c c min bsfc = 270 g/kWh (SI) 200 g/kWh (CI) Volumetric efficiency ns ma m n ma = m a air mass per cycle v = = a s 2N  aV  aV 2 N reference air mass in  aV displacement volume Air/fuel ratio A m a F/A 1 rich SI engines (gasoline) 12-18 CI engines (diesel oil) 18-100  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Relationships among parameters Thermodynamic path ( W = Q → P = Q  Q = m f qc )   m f   m f  Pb = Qb = m f qc b = m a  qc b  Pi = Qi = m f qc i = m a  qc i  m a   m a  2N 2N m a = ma = v  aV as ma = v  aV ns ns F 2N F 2N Pb = bv  aV qc Pi = iv  aV qc A ns A ns Pb F 1 Pi F 1 b = = bv  aV qc i = = iv  aV qc 2N A  ns 2N A  ns  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Relationships among parameters Mechanical path ( W = mep Ap s ) 2N Up 2N Up Pb = bmep V = bmep nc Ap Pi = imep V = imep nc Ap ns ns ns ns because V = nc Ap s U p = 2 Ns Pb 1 Pi 1 b = = bmep V i = = imep V 2N  ns 2N  ns Combining these expressions: F F F bmep =bv  a qc = i mv  a qc imep =iv  a qc A A A no reference to engine size in these formulae If bmep and mean piston speed are “known” for a good engine design, then Pb  Ap  b 2 and τb  V  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Introduction to performance curves A correct coupling between an engine and a user device involves the knowledge of engine performance curves and user device requirements: - How does engine performance vary with the operating conditions? - How can the operating conditions be defined? The variables that univocally determine engine behavior must be identified (i.e. the degrees of freedom of engine operation) Engine operation can be described according to the torques applied to the crankshaft: - indicated torque, generated by gas pressure inside the cylinder - friction torque, generated by passive resistances internal to the engine - external torque, due to the external load i i = f + ext → N = const. (b = ext)  i > f + ext → N  → N = f (m f , ext (=  b )) i < f + ext → N  f ext  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Engine operation degrees of freedom Equilibrium examples engine engine engine speed speed speed minimum fully fully opening open open engine minimum engine maximum engine intermediate no load load load idle design point variable speed operation operation operation at maximum fuel flow rate In conclusion, engine operating point is defined by two parameters. These are usually: - engine rotational speed (or mean piston speed) - “load” (which has a broad meaning, it could be throttle opening, the external load, brake torque, the percentage of maximum torque or power at a given rotational speed, bmep …)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Performance curves Pi Power and torque obtained at maximum i (or imep) load at each rotational speed  i  imep  b  bmep Pb b (or bmep) Pi / b = 2 N i / b Pb Their shape is determined by the curves tan  =  b of the indicated thermal, volumetric and N  mechanical efficiencies Nmax NmaxP N i v fluid dynamic losses m higher lower gas + inertia -> heat losses backflow N N N F F imep =iv  a qc bmep =i mv  a qc A A  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Stability of torque curve The torque curve is said to be stable or unstable according to engine response to a sudden variation of the external load 1: starting point   2: final point b b i  ext ext f ext N N b = i - f > ext → N  If the curve is descending ascending the torque difference tends to decrease increase a new equilibrium point is found is NOT found and the curve is said to be stable unstable  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Coupling with user device requirements Three categories of applications can be distinguished: - the user device works at a constant speed, and requires a variable torque that has to match the its load (e.g., an electric generator) - the user device is a turbomachinery (e.g. pump, fan, compressor, propeller), so the external load increases with the square of the rotational speed - the user device requires variable loads and rotational speeds, with rapid transients and stable operation (e.g. driving wheeled vehicles: ext = k1+k2N2 ) 2N Pb = bmep V P ns P   N N “desirable” for drive “theoretical” i.c.e. curves  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE b Performance curves for drive   I 1 ext ext w II N 2 III 1 IV 2 b b N N acc Nw  b N = w Nw “desirable” curve:   N  - all points are stable  w(i ) =  b   - max accelerating torque -> rapid transients   N w i i = I, II, III, IV  “real” i.c.e. curve: N  Nw  = N   - points at low speed are unstable  w(i )  N i - load can be higher than max torque - longer transients Pb → a gearbox is needed to “scale” torque curve Pb b b Engine elasticity: the regimes of max power and max torque should be distant to have good acceleration at low speeds Nmax NmaxP N Nmax NmaxP N  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Vehicle dynamics  Cd   a  Av U v2  Pv =  Cr  mv  g +  U v + mv  g U v sin  = k1U v + k2U v 3   2  acceleration Ww>Wv steady speed Ww=Wv deceleration Ww volumetric air/fuel ratio -> efficiencies are not efficiency is penalized, but brake significantly affected thermal efficiency is also lowered by the modifications to the indicated cycle F bmep =bv  a qc A full load b,max b,max N [rpm] N [rpm]  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Specific fuel consumption Pb [HP] > F/A SI engines bmep [bar] < i < m min bsfc (> heat loss) (> N) N [rpm] < m Pb [HP] (< load) bmep [bar] CI engines N [rpm]  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Specific fuel consumption and air/fuel ratio SI engines: A) too rich mixture -> combustion tends to be slow and unstable -> less indicated work -> high bsfc with unsufficient yield (bmep less than max) B) slightly rich mixture -> all oxygen is consumed -> maximum bmep C) stoichiometric mixture -> not all oxygen is consumed -> bmep is lower D) slightly lean mixture -> best thermal efficiency = minimum bsfc E) too lean mixture -> combustion is slow and unstable -> fuel is badly used and then bsfc increases Throttle must be partially closed to get lower loads N=const. As throttle is closed, mechanical efficiency throttle opening decreases because load decreases and pumping work is higher as well bsfc CI engines: F) maximum bmep with acceptable emissions G) to H) best bsfc -> better combustion with a higher air/fuel ratio I) minimum load -> bsfc is higher due to the lower mechanical efficiency % bmep max  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Parameters affecting engine power How can the power of a given engine be increased? 2N Up F Pb = bmep V = bmep nc Ap bmep =bv  a qc ns ns A 1) improving the brake thermal efficiency 2) improving the volumetric efficiency (ducts, valve size, timing and lift) 3) increasing air density (-> supercharging) 4) increasing F/A (but clean and efficient combustion is a constraint) 5) increasing the mean piston speed (engine rotational speed) Power is also sensible to atmospheric conditions (air density changes): p − pw T0 Pb = Pb 0 p0 T 0 = dry air in standard atmospheric conditions (1.013 bar, 288.15 K) pw = partial pressure of water vapor in the air  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Parameters affecting engine power Cylinder geometry: stroke/bore ratio lower Up, >Ap lower heat losses >v, >N, >P/V s/b < 1 = s/b > 1 >b Cylinder geometry: number of cylinders + + + = (+) lower Up -> >N, >P/V (-) longer crankshaft (torsional stress) (+) r can be raised -> >b (-) more friction surfaces -> < m (+) lower heat flow per unit area (-) higher costs for manufacturing and to be removed by cooling maintenance (+) instant torque is more uniform (+) better mass balancing -> less vibrations  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Typical design and operation data application N b s/b Up r bmep b P/V mass/P bsfc (103rpm) (mm) (m/s) (bar) (kW/dm3) (kg/kW) (g/kWh) motorcycles: SI 2s 7-10 45-70 0.8-1.0 16-20 10-15* 7-10 25-30 100-200 0.8-1.5 350 SI 4s 6-10 50-80 0.7-0.9 15-18 9-11 9-11 30-35 70-100 1-2 300 mobile equip.: SI 2s 6-8 40-70 0.8-1.0 14-18 8-12* 6-8 20-30 60-100 1-1.5 300 CI 4s 4-5 70-90 0.9-1.1 10-14 18-21 7-9 30-40 30-50 3-6 270 passenger cars: SI 4s 5.5-6.5 70-100 0.8-1.0 11-16 9-11 8-12 30-40 40-70 1-2 270 CI 4s (TC) 4-5 80-100 1.0-1.1 11-13 20-23 10-16 35-45 25-35 3-4 230 trucks: CI 4s NA 2-3 90-130 1.0-1.2 9-13 17-20 7-9 40-45 16-20 4-8 210 CI 4s TC 2-2.5 90-140 1.1-1.3 9-13 16-17 14-18 45-50 20-25 3-6 200 medium speed: CI 4s TC 0.6-1.8 150-400 1.1-1.4 8-11 15-16 15-22 45-55 10-22 8-15 190 large engines: CI 2s TC 0.08-0.2 550-850 2.0-3.6 6-8 13-15 16-18 50-60 2-5 20-40 180 LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE F7020T Internal Combustion Engines O4 – Volumetric efficiency  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Air intake process Engine performance strictly depends on air mass flow rate. If more air is brought inside the cylinder, more fuel can be burnt -> more thermal energy is released -> the area of the indicated cycle is larger -> brake torque is higher Air intake process limits engine maximum brake power. It depends on how quick: - the highest possible amount of air is aspirated - air and fuel are mixed - the mixture is burnt correctly - combustion products are expelled after being exploited as fully as possible to generate mechanical power Gas flow in intake and exhaust systems is unsteady (pulsating). The energy involved can be exploited to help the gas exchange process if the systems are properly designed (tuned) Gas exchange process differs in 4-stroke and 2-stroke engines: 4s) two stokes are spent for it and piston plays an active role -> efficient process 2s) the process occurs when piston is near BDC, just because fresh charge has been pressurized -> less efficient process  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Air intake process (4-stroke engines) Air intake process strictly depends on pressure history in the intake system and inside the cylinder This history is difficult to be evaluted analytically (too many variables and effects) An index is used as an overall measure of the effectiveness of the process: volumetric efficiency  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Volumetric efficiency Compares the actual mass of aspirated air per operating cycle with a reference amount of air, i.e. the mass of air that would occupy the displacement volume at atmospheric conditions ma m a v = =  a ,0V  V N 2 a ,0 ns This definition is valid for both SI and CI engines, and for any fuel feeding system This index is not really an efficiency, in particular cases it can be greater than one (ram effects)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Variables and effects Design and operating variables affecting volumetric efficiency are: - fuel type, equivalence ratio, fuel evaporation in the intake system, fuel latent heat - air temperature variation in the intake system - pressure in intake and exhaust systems - compression ratio - engine rotational speed - design of ducts, manifolds and ports of intake and exhaust systems - geometry, size, maximum lift and timing of intake and exhaust valves These variables produce effects, which can be - independent of piston motion - quasi-static (dependent on mean piston speed only) - dynamic (dependent on the pulsating flow within intake and exhaust systems)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Fuel and equivalence ratio In SI engines fuel and water vapor reduce the partial pressure of air in the aspirated mixture. According to ideal gas law The simpler the molecule of fuel and the higher the equivalence ratio (with the vaporized fraction of the fuel), the lower the volumetric efficiency  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Fuel latent heat In SI engines the evaporation of fuel in the intake ducts reduces the temperature of the mixture -> air density is higher -> volumetric efficiency is improved Enthalpy balance before (B) and after (A) evaporation of fuel mass fraction xe ( Q + m a ha + m f h fL )B = (m a ha + (1− xe )m f h fL + xe m f h fV )A (Q / m a ) − xe ( F / A)h fLV TA − TB = c p ,a + ( F / A)c fL Q = heat provided to the mixture h = cT L = liquid c p ,a = air specific heat at constant pressure V = vaporized c fL = specific heat of the liquid fuel If xe=1 and no heat is provided, then the temperature drop would be significant at stoichiometric equivalence ratios (–19°C for isooctane and -128°C for methanol) In general, heat is always provided to ensure complete evaporation  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Inlet and exhaust pressure ratio and compression ratio When intake pressure is lower than exhaust pressure (usual case), volumetric efficiency increases with the compression ratio (a lower mass of residual gas is trapped in the clearance volume). If intake pressure is higher than exhaust pressure, volumetric efficiency decreases with the compression ratio (in this case the clearance volume is filled by the fresh charge)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Heat transfer in the intake system Intake system ducts are hotter than the fresh charge, so its temperature tends to increase. The intake system can be considered as a simple duct with an equivalent diameter (de) and length (L), wall temperature being Tw Tw Ta = f (U p , L, d e ,Tw ) Volumetric efficiency decreases as: - mean piston speed is decreased (air velocity decreases and the residence time of the fresh charge in the duct is increased) - L is increased (heat transfer area increases) - de is decreased (surface/volume ratio increases for a unit of mass of the fresh charge -> more heat transfer) - Tw increases (it is a function of engine thermal regime, that is of equivalence ratio and coolant temperature)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Friction losses Each part of the intake system generates concentrated and/or distributed friction losses which are proportional to the square of air velocity in that part p pi 2 vi = =  i  i  i 2 Since air velocity in the intake system is proportional to mean piston speed, the pressure drop is proportional to the square of mean piston speed 2 p U p  = 2  'i i  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve timing Valve timing does not coincide with top and bottom dead centers Intake valve opens before the TDC to be wide open when the intake stroke begins Intake valve closes well after the BDC to exploit the inertia of the fresh charge entering the cylinder, so that it continues to fill the cylinder when the compression stroke has started Exhaust valve opens well before the BDC to perform the blowdown phase of the exhaust process, so that cylinder pressure is close to ambient pressure when the displacement phase (i.e. the exhaust stroke) begins Exhaust valve closes after the TDC to exploit the inertia of the residual gas exiting the cylinder, so that more residual gas can be extracted from the clearance volume Both valves are open during the overlap period. Since exahust pressure is usually higher than intake pressure, backflow of residual gas in the intake port occurs if the inertia of the residual gas is not sufficient (flow areas are small)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve timing Volumetric efficiency is affected by intake valve lag and overlap period: at high speeds both must be increased to fill the cylinder effectively, but this results in backflow at low speeds.  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Port and valve size. Valve lift Increasing the size of valves and ports increases the volumetric efficiency at high speeds (lower friction losses), but reduces the inertia of the fluid columns -> more backflows at low speeds Valve lift cannot be increased beyond a critical value (port area becomes the bottleneck)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Pressure waves (tuning) Pressure waves propagate at the local sound speed in the intake and exhaust system, and they can help or inhibit the gas exchange process The unsteady exhaust process causes compression waves to be propagated forward in the exhaust system. If the reflected pressure wave reaches the cylinder at the end of the exhaust process, decreasing the pressure in the exhaust port, it helps the extraction of the residual gas, and the system is said to be tuned The unsteady intake process causes expansion waves to be propagated backward in the intake system. If the reflected pressure wave reaches the cylinder at the end of the intake process, increasing the pressure in the intake port, it helps filling the cylinder with fresh charge, and the system is said to be tuned Tuning cannot be obtained at all engine speeds: it naturally depends on the frequency of the intake/exhaust processes (that is on engine speed) and on the geometry (lengths and volumes) of the intake/exhaust systems. If suitable volumes and lengths are chosen, a resonance frequency can be selected in order to exploit pressure waves to improve the volumetric efficiency in a particular range of engine speed  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Pressure waves (tuning)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Summary of the effects on volumetric efficiency LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE F7020T Internal Combustion Engines O5 – Valves  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Introduction In 4-stroke engines the poppet valves are used to control intake and exhaust flow areas. Valves and ports together are the most important flow restrictions in the intake and exhaust flow system. The characteristics of flow through poppet valves are critical because high pressure drops are generated -> pumping work is increased and volumetric efficiency is reduced. The pressure drop depends on the square of mean piston speed and on flow area (which should be as large as possible) For the intake system (with fully open throttle, if present): 2 Up patm − pcyl  (0.03  0.12) patm  (20  80)  atm (2500-10000 Pa) 2 and the inlet valve is responsible for the 50-70% of the overall pressure drop For the exhaust system: 2 Up pcyl − patm  (0.05  0.20) patm  (40 150)  atm 2  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve and port geometry Valve head is used to control minimum flow area (diameter) Valve stem is used to drive valve movement and to transmit the command given by the camshaft (diameter) Valve seat has a conical shape to avoid gas leakage when cylinder pressure is high (inner diameter, width, angle (30° or 45°)) Valves are opened by making them slide along stem axis towards the piston (lift)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve and port geometry Design recommendations: - cross-sectional areas should be as large as possible to increase volumetric efficiency (and then power) at high engine speeds (but this causes backflow and lower power at low engine speeds) - the port increases the friction losses due to the valve alone (largest possible curvature radius, minimum protrusion of valve guide) - intake valves are slightly larger than exhaust valves (higher volumetric efficiency and exhaust valve heat transfer issues through the seat and the guide) INTAKE EXHAUST  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve size and cylinder bore Ideal case d v max = B d v max = 1 B (B = (1+ 2 )d ) v max 2 1+ 2 Av max Av max = 0.5 2 = 0.25 = 2 (0.41) 2 = 0.34 (+30%) Ac Ac Real case chamber flat wedge hemispherical 4 valves Dv (intake) (0.42-0.44)B (0.43-0.46)B (0.48-0.50)B (0.35-0.37)B Dv (exhaust) (0.34-0.37)B (0.35-0.37)B (0.41-0.43)B (0.28-0.32)B  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Number of valves Intake flow area can be increased in two ways only: - increasing the number of valves to better exploit cylinder section - inclining valve axes in a domed head surface  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve timing Valve motion cannot occur instantaneously: the laws of dynamics must be respected - EVO 40°-60° BBDC, as a compromise between the work lost at the end of the expansion stroke and the work spent during the exhaust stroke - EVC 10°-30° ATDC to exploit gas inertia during the overlap phase - IVO 10°-40° BTDC to be sufficiently open at TDC: this contributes to gas exchange in the clearance volume - IVC 40°-80° ABDC to make fresh charge enter the cylinder thanks to its inertia  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve timing Inlet valve closing lag Has a major influence on the [rev/min] shape of volumetric efficiency inlet valve lag volumetric efficiency curve (compromise between backflow at low speeds and inlet valve lag unexploited inertia at high speeds) Overlap period engine speed [rev/s] Two issues: 1) unburnt charge flows directly in the exhaust port (EV lag too long at low speeds) -> UHC and CO emissions 2) residual gas flows back in the intake port (IV lead too large at partial throttle) -> combustion is slower -> UHC and CO emissions  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve lift Valve motion is commanded by the camshaft (one rev per 2 crankshaft rev) It is designed according as a compromise among: - maximum volumetric efficiency - no oscillatory motions superimposed to cam motion (vibrations or impacts) - low mechanical friction between surfaces in contact Valve lift controls minimum flow area (until port area becomes the bottleneck) 1 2 3 ≈ 0.25  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Variable valve timing In traditional engines both valve timing and lift curve are fixed Performance and thermal efficiency can be improved by controlling timing and lift curve as a function of load and speed 1) shifted timing angles with fixed duration and lift 2) discrete control of timing and lift 3) complete flexibility in timing and lift control 1) 2) 3)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve flow Av p0,T0 U GAS DYNAMICS – IDEAL CASE pv Hypotheses: steady, one-dimensional, isentropic and adiabatic flow of a perfect gas in a converging nozzle emptying in a large volume With reference to the minimum cross sectional area of the nozzle (Av), 1 the mass flow rate is   +1  2  2  2   pv    pv    m id =  0 Av RT0    −    −1   p0   p0       until choking occurs at the critical pressure ratio pv  pv  2   −1 =   = 1/1.86 = 0.538 for  =1.35 p0   + 1  pv and the choked mass flow rate is  +1  2  2( −1) m id ,choked =  0 Av RT0     + 1   LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Valve flow REAL CASE Real flows do not satisfy the hypotheses of the ideal case: - specific heat ratio varies with temperature (no perfect gas) - velocity distribution is not uniform in a cross sectional area (no 1D flow) - friction and heat transfer occur (no isentropic, no adiabatic) A coefficient C is used to take into account the difference between the actual mass flow rate and the ideal mass flow rate 1   2  +1  2  +1  2   pv    pv     2  2( −1) m = Cm id = CAv  0 RT0    −   or CAv  0 RT0    −1  p0   p0     + 1       For similar geometry C=f(Re,Ma). For Ma no - valve head area: it is constant -> the flow coefficient Cf is simply proportional to the actual mass flow rate:  Aref = d 2 m = C f m id ( Aref , p0 ,T0 , pv ) = C f  const. 4 It comprises the effects due to both the variation of flow area and the real fluid-dynamic conditions. In order to focus on fluid-dynamic details only, a variable area must be chosen -> - valve curtain area: the cylindrical surface defined by valve head diameter and lift, varies linearly with lift -> the discharge coefficient Cd is more suitable to show the effectiveness in using the available flow area Aref =  d l m = Cd m id ( Aref , p0 ,T0 , pv )  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Intake valve discharge coefficient Three flow regimes can be identified in the diagram, due to: - inertia: the flow tends to maintain its direction and separates from the wall -> jet - viscosity: the jet tends to involve the surrounding fluid in a recirculation which generates a depression -> the flow may reattach to the wall Cd is a measure of the effectiveness in exploiting the available curtain area -> lower Cd when the flow is separated and/or the boundary layer is relatively thick Seat should be as small as possible with =30° a) low lift b) medium lift c) high lift Cd Cd Cd seat angle l/Dv l/Dv l/Dv  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Intake valve discharge coefficient Rounding the upstream corner of the valve seat reduces the tendency of the flow to separate -> higher Cd is obtained The port does not alter the Cd if the design has followed the recommendations mentioned above: - a large curvature radius is used - the cross sectional area near the protrusion of the valve is increased These recommendations may not be followed if organized charge motions within the cylinder have to be generated (but the resulting Cd is lower) port 1 port 2 port 3 valve alone Cd Cd seat l/Dv l/Dv  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Exhaust valve discharge coefficient - Two flow regimes can be identified, a) low lift but the change is not abrupt b) high lift - Higher seat angles perform better at high lifts - The outer corner of valve seat should be rounded to decrease flow separation Cd - Port design recommendations are similar to those given for the intake port l/Dv seat valve alone valve Cd Cd seat ang. blunt corner seat ang. blunt corner seat ang. port 1 port 2 port 3 l/Dv l/Dv  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Mach index Compressibility effects on the intake discharge coefficient: if valve head diameter is not large enough, flow may choke at high speed (critical for volumetric efficiency). Volumetric efficiency can be correlated with a Mach index Z: AcylU p Z= C fi Aiv ci in which here Aiv is the area of the inlet valve(s) and Cfi is the average flow coefficient during the time in which the valve(s) is (are) open. The index is equivalent to the ratio between the velocity of an ideal volumetric flow rate volumetric efficiency (the ideal filling of the cylinder) through an average effective inlet area and the inlet sound speed (it is a Mach number) The correlation shows that this index should be kept lower than 0.6 in order to avoid choking in the normal range of engine operation Mach index Z  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Measurement of the discharge coefficient Flow and discharge coefficients are air flow measured using benches equipped meas. device with a fan that generates a steady flow. Several measures can be taken at different valve lifts and pressure air exit ratios air suction air flow meas. device LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE F7020T Internal Combustion Engines O7 – Supercharging  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Introduction Engine power can be boosted if the air entering the cylinder has a density which is higher than that at ambient conditions (cylinder supercharging). F 2N Pb = b v  a V qc A ns This means that air has been compressed and/or cooled outside the cylinder supercharged supercharged naturally aspirated naturally aspirated Performance Performance comparison between mixture temp. comparison between a mixture temp. a 2-liter supercharged 1.6-liter supercharged boost pressure boost pressure engine and engine and a 2-liter naturally a 2-liter naturally aspirated engine aspirated engine SFC SFC power power torque torque engine speed (rpm) engine speed (rpm)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Effects of supercharging If air density is higher, air mass flow rate is higher and a higher fuel mass flow rate can be burnt in the same displacement volume -> the purpose is to increase engine power per unit of volume and to reduce engine weight per unit of power Mechanical efficiency may increase (higher friction due to higher loads, but no more pumping work if intake pressure exceeds exhaust pressure; compressor power may be free or taken into account among engine auxiliaries) Volumetric efficiency may increase (gas exchange process is more effective, in particular large overlap angles can be used in CI engines) Indicated thermal efficiency may increase (relatively lower heat loss: more thermal energy is released in the same volume enclosed by the same surface) Higher mechanical and thermal loads (detonation may occur in SI engine, thermal stresses set a limit to supercharging in CI engines) Lower pollutant emissions (a lower equivalence ratio is used in CI engines)  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Cycle comparison pressure pressure Pressure in the intake stroke is above ambient pressure -> the gas exchange process may contribute to the positive volume indicated work volume Vc Vc The higher thermal energy release can be used to reach higher pressures using naturally the same compression ratio or aspirated to “inflate” cycle area by reducing the super- pressure pressure charged pmax = const. compression ratio in order to reach the r = const. same maximum pressure pi pi pressure (bar) volume volume Real indicated cycles show a slight increase in maximum pressure but a Vc volume large increase in cycle area  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Supercharging methods Can be classified according to: - the engine (2- or 4-stroke, SI or CI) - the compressor (volumetric, mechanical supercharger turbocharger centrifugal, pressure wave) - the source of the power required to drive the compressor: ◼ mechanical supercharging (driven by crankshaft, pi > pa and pe = pa) ◼ turbocharging (driven by an exhaust gas turbine, both pi and pe > pa, can be constant-pressure or pulse) ◼pressure-wave supercharging (uses engine-driven compressor and turbocharger 2-stage turbocharger wave action in the intake and exhaust systems, pi > pa and pe=pa) Combined schemes can also be used An intercooler can be introduced after the charge has been compressed turbocharger + turbocompund with intercooler  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE General trends in 4-stroke supercharged engines Power per unit of volume increases because: - density is increased - intake pressure is usually greater than exhaust pressure -> no pumping work and positive contribution to indicated work - volumetric efficiency is increased (fresh charge in the clearance volume, in particular in CI engines with large overlap period) On the other hand, different aspects affect brake thermal efficiency: - the ideal thermal efficiency is reduced as the compression ratio is decreased (in particular in SI engines to avoid detonation) - heat losses are relatively less important, so the indicated efficiency increases - friction losses increase less than imep, so mechanical efficiency should increase - mechanical supercharging makes mechanical efficiency drop significantly There are two clear trends: - SI engines with mechanical supercharging will have a lower brake thermal efficiency due to the lower compression ratio and the lower mechanical efficiency - CI engines with turbocharging will have a higher brake thermal efficiency because the indicated thermal efficiency is not significantly reduced while the mechanical efficiency is considerably increased  LULEÅ UNIVERSITY OF TECHNOLOGY – DIVISION OF ENERGY SCIENCE Turbomachinery for supercharging 1s