2 Introduction into the functioning of internal combustion engines

2.1 Energy conversion

In energy conversion, we can distinguish hierarchically between general, thermal, and motor energy conversion.

Undergeneral energy conversion is understood the transformation of primary into secondary energy through a technical process in an energy conversion plant, see Fig. 2.1.

PrimaryenergyOilderivativesNaturalgasHydrogenBiomassWindWaterSunE.C.P.FurnaceinternalcombustionengineGasturbineFuelcellPowerstationWindwheelHydraulicturbineFotocellSecondaryenergyThermalenergyMechanicalenergyElectricenergyElectricenergyFig. 2.1:Diagram of general energy conversionThermal energy conversion is subject to the laws of thermodynamics and can be described formally, as is shown in Fig. 2.2. .QsuppliedThermalenergyPtconversionplant.QremovedFirstlawofthermodynamics:..Pt=Qsuppl.-Qremov.Fromthesecondlawofthermodynamicsfollows:.Qremov.>0!Thermalefficiency:.PtQremov.=1-.<1hth=.Qsuppl.Qsuppl.Fig. 2.2: Diagram of thermal energy conversionTheinternal combustion engine and the gas turbine are specialized energy conversion plants, in which the chemical energy bound in the fuel is at first transformed into thermal energy in 62 Introduction into the functioning of internal combustion engines

the combustion space or chamber, this being then transformed into mechanical energy by the motor. In the case of the stationary gas turbine plant, the mechanical energy is then converted into electrical energy by the secondary generator.

Chemical energybound in fuelCombustion processThermal energyDriving mechanismMechanical energyGeneratorElectrical energyInternalcombustionengineGasturbineFig. 2.3: Diagram of energy conversion in an internal combustion engine or gas turbine

2.2 Reciprocating engines

Internal combustion engines are piston machines, whereby one distinguishes, according to the design of the combustion space or the pistons, between reciprocating engines and rotary en-gines with a rotating piston movement. Fig. 2.4 shows principle sketches of possible struc-tural shapes of reciprocating engines, whereby today only variants 1, 2, and 4 are, practically speaking, still being built.

1352461In-lineengine2V-engine3Radialengine4FlatengineMulti-pistonunits:5Dual-pistonengine6OpposedpistonengineFig. 2.4: Types of reciprocating engines

2.2 Reciprocating engines 7

For an extensive description of other models of the combustion engine, see Basshuysen and Schäfer (2003) and Maas (1979).

2.2.1 The crankshaft drive

The motor transforms the oscillating movement of the piston into the rotating movement of the crankshaft, see Fig. 2.5. The piston reverses its movement at the top dead center (TDC) and at the bottom dead center (BDC). At both of these dead point positions, the speed of the piston is equal to zero, whilst the acceleration is at the maximum. Between the top dead cen-ter and the underside of the cylinder head, the compression volume Vc remains (also the so-called dead space in the case of reciprocating compressors).

Intakeair+fuelTDCCombustionchamberPistonringsBDCDischargeexhaustgasCylinderheadCylinderlinerPistonPistonpinConnectingrodCrankcase(qualitative)IntegratedcounterweightCrankpinBasepinCrankshaftFig. 2.5: Assembly of the reciprocating engine Fig. 2.6 shows the kinematics of a crankshaft drive with crossing, in which the longitudinal crankshaft axle does not intersect with the longitudinal cylinder axle, but rather is displacedby the length e.82 Introduction into the functioning of internal combustion engines

sComp.s(j)lc2c3brjc1e>0Fig. 2.6: Kinematics of the crankshaft drive For the piston path s(ϕ), it follows from Fig. 2.6: s(ϕ)=c3−c2−rcos(ϕ−β)from which with sinβ=e§e·r+landβ=arcsin©¨¨r+l¸¸¹,respectivelyc1=e−rsin(β−ϕ),c2=l2−c12andc3=(r+l)2−e2finallys(ϕ)=(r+l)2−e2−l2−[e+rsin(ϕ−β)]2−rcos(ϕ−β)results. The derivative provides for the piston speed the relation dsdϕ=rsin(ϕ−β)+r[e+rsin(ϕ−β)]cos(ϕ−β)l2−[e+rsin(ϕ−β)2] . With the definition of the cylinder volume (2.1)(2.2)(2.3)2.2 Reciprocating engines 9

V(ϕ)=Vc+D2

π4

s(ϕ)(2.4)

follows for the alteration of cylinder volume

πdsdV

. =D2

dϕ4dϕWith the eccentric rod relation λe=rl, it follows finally for the limiting case e=0

­1ªº½22s(ϕ)=r®[1−cos(ϕ)]+1−1−λsin(ϕ)¾e»λe«¼¿¬¯

(2.5)

(2.6)

and

ª

λds

=r«sin(ϕ)+e

«dϕ2«¬

º

» .»2

1−λ2esin(ϕ)»¼sin(2ϕ)

(2.7)

2.2.2 Gas and inertia forces

The motor is driven by the gas pressure p(ϕ) present in the combustion space. With the

piston area AP=D2π4, one then obtains for the gas force

Fg=D2

π4

p(ϕ) . (2.8)

Because of the masses in motion in the driving mechanism, additional and temporally vari-able inertia forces arise, which lead to rotating and oscillating unbalances and must at least partially be counterbalanced in order to guarantee the required driving mechanism running smoothness. The single components of the motor execute rotating (crank pin, mc.p.), oscillat-ing (piston block, mP), or mixed (connecting rod) movements. If one distributes the mass of the connecting rod into a rotating (mc.r.,rot) and an oscillating (mc.r.,osc) portion, one then obtains for the rotating and oscillating masses of the driving mechanism

mrot=mc.r.,rot+mc.p.,mosc=mc.r.,osc+mp.

For small λe, the expression under the root in (2.6) corresponding to

1−

λ2e

sin(ϕ)=1−

2

2

λe

2

sin(ϕ)−

2

λ4e

8

sin4(ϕ)−...

can be developed into a Taylor’s series, whereby the third term for λe=0.25 already be-comes smaller than 0.00048 and can thus be neglected as a rule. With the help of trigonomet-ric transformations, one finally obtains for the piston path

λs

=1−cos(ϕ)+e(1−cos(2ϕ)) . r4

With the angular velocity ω(2.9)

102 Introduction into the functioning of internal combustion engines

dϕ=ωdt

one obtains for the piston speed

dsdsdϕds==ωdtdϕdtdϕthe expression

λdsªº=rω«sin(ϕ)+esin(2ϕ)»dt2¬¼

and for the piston acceleration

d2sdt2

2

d2s§dϕ·2ds=¨¸=ωdϕ2©dt¹dϕ2

2

(2.10)

finally

d2sdt2

=rω2[cos(ϕ)+λecos(2ϕ)] .

(2.11)

With that, one obtains the expression

Fm,rot=mrotrω2

(2.12)

for the rotating inertia force, which triggers an unbalance striking in the crankshaft axle and

rotating with the speed of the crankshaft. For the oscillating inertia force one obtains the expression

Fm,osc=moscrω2[cos(ϕ)+λecos(2ϕ)] .

(2.13)

This consists of two parts, whereby the first rotates with simple crankshaft speed and the second with doubled crankshaft speed. One therefore distinguishes between inertia forces of first and second order,

F1=moscrω2cos(ϕ),F2=moscrω2λecos(2ω).

The inertia forces are proportional to ω2 and are thus strongly contingent on speed. The resulting piston force consists of gas force and oscillating inertia force,

FP=D2

π4

p(ϕ)+moscrω2[cos(ϕ)+λecos(2ϕ)] .

(2.14)

Fig. 2.7 shows the progression of gas force and oscillating inertia force for a 4-stroke recipro-cating engine over a full cycle. One recognizes that the peak load caused by the gas force is quickly diminished with rising speed because of the inertia force proportional to ω2.

2.2 Reciprocating engines 11

PressureFgFm0FgFm01803600j[°CA]720Fig. 2.7: Gas force and oscillating inertia force of a 4-stroke reciprocating engine 2.2.3 Procedure With regard to charge changing in the reciprocating engine, one distinguishes between the 4-stroke and the 2-stroke methods and in reference to the combustion process between diesel and spark-ignition (SI) engines. In the case of the 4-stroke-procedure, see also Fig. 2.8, the charge changing occurs in both strokes, expulsion and intake, which is governed by the dis-placement effect of the piston and by the valves. The intake and exhaust valves open before and close after the dead point positions, whereby an early opening of the exhaust valve indeed leads to losses during expansion, but also leads to a diminishment of expulsion work. With increasing valve intersection, the scavenging losses increase, and the operative efficiency decreases. Modern 4-stroke engines are equipped, as a rule, with two intake and two exhaust valves.ppzppzEVOpuVcIVOEVCVdEVOIVCVpuVcVdIVCEVCIVO2-Stroke-Cycle processV4-Stroke-Cycle processFig. 2.8:p,V diagram for the 4-stroke and 2-stroke processes In the case of the 2-stroke engine, the charge changing occurs while the piston is near the BDC. With such so-called piston-valve engines, the exhaust gas is expelled out of the cylin-122 Introduction into the functioning of internal combustion engines

der by the in-flowing fresh air, if the piston sweeps over the intake and exhaust sections ar-ranged in the lower area of the cylinder. In the case of larger engines, exhaust valves are mostly used instead of exhaust ports, which are then housed in the cylinder head. Instead of so-called loop scavenging, one then has the fundamentally more effective uniflow scaveng-ing. For more details, see Merker and Gerstle (1997).

2.3 Thermodynamics of the internal combustion engine

2.3.1 Foundations

Our goal in this chapter will be to explain the basic foundations of thermodynamics without going into excessive detail. Extensive presentations can be found in Baehr (2000), Hahne (2000), Lucas (2001), and Stephan and Mayinger (1998, 1999).

For the simulation of combustion-engine processes, the internal combustion engine is sepa-rated into single components or partial systems, which one can principally view either as closed or open thermodynamic systems. For the balancing of these systems, one uses the mass balance (equation of continuity)

dm

󰀆1−m󰀆2=mdt

(2.15)

and the energy balance (1st law of thermodynamics)

dU󰀆+W󰀆+E󰀆+E󰀆=Q12dt

(2.16)

with

§c2·󰀆󰀆¨E=m¸¨h+2¸

¹©

for the open, stationary flooded system shown in Fig. 2.9 (flow system), or

dU󰀆+W󰀆=Q

dt

(2.17)

for the closed system shown in Fig. 2.10 (combustion chamber).

.Q.m1p1T1c1A1h1.m2p2T2c2A2h2U.WFig. 2.9: Open thermodynamic system (----- system boundaries) 2.3 Thermodynamics of the internal combustion engine 13

.QU,m,VT,p.WFig. 2.10: Closed thermodynamic system (----- system boundaries)In closed systems, no mass, and with that no enthalpy, flows over the system limits. Neglect-ing the blow-by losses, the combustion chamber (cylinder) can be viewed as a closed system during the so-called high pressure process (compression and expansion act). In contrast, in the case of an open system, e.g. a reservoir or a line section, masses can flow over the system boundaries.Neglecting the friction or dissipation of mechanical work into heat, one obtains for the vol-ume work󰀆=−pdV . Wdt(2.18)In the open system, one summarizes the thermal energy transferred to the system boundaries and the intake and expulsion work practically as enthalpy h≡u+pv . (2.19)(2.20)Thethermal state equationf(p,T,v)=0ties together the three thermal condition magnitudes of pressure, temperature, and volume and thecaloric state equationu=u(T,v)andh=h(p,T),respectively(2.21)describes the inner energy as a function of temperature and volume, or the enthalpy as a func-tion of pressure and temperature. We will in the following view the materials under considera-tion first as ideal gases, for which the thermal state equation pv=RT(2.22)is applicable. Because the inner energy of ideal gas is only dependent on temperature, follows from (2.19) with (2.22), that this is also valid for enthalpy. Thus for differential alteration of caloric magnitudes of the ideal gas we have: du=cv(T)dTanddh=cp(T)dT,respectively.(2.23)142 Introduction into the functioning of internal combustion engines

For ideal gas

R=cp(T)−cv(T)(2.24)

and

κ=

cpcv

(2.25)

are applicable. For reversible condition alterations, the 2nd law of thermodynamics holds in the form

Tds=dq .

(2.26)(2.27)

With that, it follows from (2.17) with (2.18)

du=−pdv+Tds .

With (2.23), it follows for the rise of the isochores of a ideal gas

T§dT·

. ¨¸=

dsc©¹sv

(2.28)

Analogous to this, it follows for the rise of isobars

T§dT·

, ¨¸=

©ds¹scp

and for the isotherms and isentropes follows

dpdppp=−κ . =− or

dvvvdv

Fig. 2.11 shows the progression of simple state changes in the p,v and T,s diagram.

pIsothermIsentropeIsochoreIsobarTIsochoreIsobarIsothermIsentropevFig. 2.11: Course of a simple change of state in the p, v- and in the T, s diagram sWith the relations above, one finally obtains for the energy balance of the closed system2.3 Thermodynamics of the internal combustion engine 15

mcv

dTdQdv

. =−p

dtdtdt

(2.29)

Under consideration of the enthalpy flows and the transferred kinetic energy to the system

boundaries, one obtains for the energy balance of the open system

2·2·§§c1c2dTdmdQdW¸¸ . ¨¨󰀆1h1+󰀆2h2+mcv+cvT=++m−m

¨¨dtdtdtdt2¸2¸¹¹©©

(2.30)

For stationarily flooded open systems, it follows for the case that no work is transferred

22·ºª§c2c1

¸»=dQ . ¨󰀆«(h2−h1)+−m

¨22¸dt«¹»©¬¼

(2.31)

With this relation, the flow or outflow equation for the calculation of the mass flows through

throttle locations or valves can be derived. We consider an outflow process from an infinitely large reservoir and presume that the flow proceeds adiabatically. With the indices \"0\" for the interior of the reservoir and \"1\" for the outflow cross section, it follows with c0=0 from (2.31)

2c1

=h0−h1 . 2

(2.32)

With the adiabatic relation

T1

¸=¨¸T0¨p©0¹

κ−1§p1·κ(2.33)

it first follows

2c1

ª

§«T1·

¸=cpT0«1−1=cpT0¨−¨T0¸2¹©«

¬

κ−1º§p1·κ»

κ−1º§p1·κ»

¸¨¨p¸©0¹

»»¼

(2.34)

and furthermore for the velocity c1 in the outflow cross section

c1=

ª«2κRT0«1−κ−1«

¬

1

¸¨¨p¸©0¹

» . »¼

(2.35)

With the equation for ideal gas, it follows for the density ratio from (2.33)

ρ1§p1·κ¸ . =¨¸ρ0¨p©0¹

With this results for the mass flow

(2.36)

162 Introduction into the functioning of internal combustion engines

󰀆=A1ρ1c1m

in the outflow cross section the relation

§p1·

¸󰀆=A1ρ0p0Ψ¨m¨p,κ¸ ,

©0¹

(2.37)

whereby

§p1·

¸Ψ¨¨p,κ¸=

©0¹

κ+1º2ª

§p1·κ»2ꫧp1·κ¸¸¨¨−«»¸¨p¸κ−1«¨p©0¹©0¹»

¬¼

(2.38)

is the so-called outflow function, which is solely contingent upon the pressure ratio p1p0

and from the isentrope exponent κ. Fig. 2.12 shows the progression of the outflow function for various isentrope exponents.

0.5Yk=1.41.31.2p*p00.40.30.20.10.00.00.20.40.60.8pp01.0Fig. 2.12: Outflow function Ψ¨¨§p1©p0

,κ¸¸·¹The maximums of the outflow function result from the relation ∂Ψ§p1∂¨¨p©0

·¸¸¹=0forΨ=Ψmax . (2.39)2.3 Thermodynamics of the internal combustion engine 17

With this, one obtains for the so-called critical pressure ratio the relation

§p1·§2·κ¸¨¨=¨κ+1¸¸¨p¸

¹©0¹crit©c1,crit=κRT1

κ−1

or

§T1

¨¨T©0·2¸ . =¸

¹critκ+1

(2.40)

If we put this relation into (2.35) for the isentropic outflow velocity, then

(2.41)

finally follows. From (2.36) follows

dpp

=κ=κRT

ρdρ(2.42)

for the isentropic flow. With the definition of sound speed

a≡

dpdρ(2.43)

thus follows

a1=κRT1

(2.44)

for the velocity of the outflow cross section. The flow velocity in the narrowest cross section of a throttle location or in the valve can thereby reach maximal sonic speed.

2.3.2 Closed cycles

The simplest models for the actual engine process are closed, internally reversible cycles with heat supply and removal, which are characterized by the following properties:

- - - •

the chemical transformation of fuel as a result of combustion are replaced by a corre-sponding heat supply,

the charge changing process is replaced by a corresponding heat removal air, seen as a ideal gas, is chosen as a working medium.

The Carnot cycle

The Carnot cycle, represented in Fig. 2.13, is the cycle with the highest thermal efficiency and thus the ideal process. Heat supply results from a heat bath of temperature T3, heat re-moval to a heat bath with temperature T1. The compression of 2→3 and 4→1 always takes place isentropically. With the thermal efficiency

qsuppliedqremoved

ηth=1−

we obtain the well-known relation

182 Introduction into the functioning of internal combustion engines

ηth,c=1−

§T1·T1

¸=f¨¨T¸T3©3¹

(2.45)

for the Carnot cycle.

qsupplp3=pmax34pT3p1=pu214qremoveds21vFig. 2.13: Carnot cycle

The Carnot cycle cannot however be realized in internal combustion engines, because

- -

the isothermal expansion with qsupplied at T3=const. and the isothermal compres-sion with qremoved at T1=const. are not practically feasible, and

the surface in the p,v diagram and thus the internal work is extremely small even at

high pressure ratios.

w

.

v1−v3

p3

,p4p2p1

In accordance with the definition, for the medium pressure of the process is applicable

pm=

(2.46)

For the supplied and removed heat amounts in isothermal compression and expansion

qsupplied=q34=RT3lnqremoved=q12=RT1ln

2.3 Thermodynamics of the internal combustion engine 19

applies. With the thermally and calorically ideal gas, we obtain

p3§T3·κ−1p4§T4·κ−1

¸=¨ and =¨¸¨T¸¸p2¨Tp1©2¹©1¹

κκfor the isentrope, from which follows

p3ppp=4 and 3=2p2p1p4p1

becauseT1=T2 and T3=T4. At first we obtain for the medium pressure

R(T3−T1)lnpm=

v1−v3p3

p1

p3p4

and finally by means of simple conversion

pm

=p1

§T3·§p3T3·κ¸¸¨¨−−1lnln¨T¸¨p¸Tκ−111¹¹©©1

.

p3T3

−p1T1

(2.47)

The relation

§T3p3·pm

=f¨¨T,p,κ¸¸p1©11¹

setting to zero and with extreme values is graphically presented in Fig. 2.14 for κ=1,4.

4pmp130.8hthp3200=p1150100hth0.620.4502510.2001234T3T15Fig. 2.14: Medium pressure of the Carnot cycle 202 Introduction into the functioning of internal combustion engines

While the thermal efficiency in an optimally run process at a pressure ratio of 200 with 0.6 achieves relatively high values, the reachable medium pressure still amounts only to pm=3.18p1. The work to be gained is thus so small that an engine realizing the Carnot cycle could in the best scenario overcome internal friction and can therefore deliver practi-cally no performance.

The Carnot cycle is thus only of interest as a theoretical comparative process. In this context, we can only point out to its fundamental importance in connection to energy considerations. •

The constant-volume process

The constant-volume process is thermodynamically efficient and, in principle, feasible cycle (see Fig. 2.15). In contrast to the Carnot process, it avoids isothermal expansion and compres-sion and the unrealistically high pressure ratio. It consists of two isentropes and two isocho-res.

p3pmaxTTmaxqsuppl.3qsuppl.421TDCBDC4qremov.v21qremov.sFig. 2.15: Representation of the constant-volume cycle in the p,v and T,s diagram It is called the constant-volume process because the heat supply (instead of combustion) en-sues in constant space, i.e. under constant volume. Because the piston moves continuously, the heat supply would have to occur infinitely fast, i.e. abruptly. However, that is not realisti-cally feasible. For the thermal efficiency of this process follows c(T−T1)qremovedT=1−1=1−v4T2cv(T3−T2)qsuppliedT4−1T1 . T3−1T2ηth,v=1−With the relations for the adiabatic T1§v1·¸¨¸=¨v2¹T2©T3§v4·¨¸=¨¸v3¹T4©κ−1κ−1§v¨1=¨©v2·¸¸¹κ½°°T°T4=3=¾−1T2°T1°°¿2.3 Thermodynamics of the internal combustion engine 21

and the compression ratio ε=v1v2 follows finally for the thermal efficiency of the con-stant-volume process

ηth,v

§1·=1−¨¸

©ε¹

κ−1

.

(2.48)

This relation, represented in Fig. 2.16, makes it clear that, after a certain compression ratio, no significant increase in the thermal efficiency is achievable.

0.8hth,V0.6k=1.2k=1.40.40.20.00148121620e24Fig. 2.16: Efficiency of the constant-volume cycle

• The Constant-pressure process

In the case of high-compressing engines, the compression pressure p2 is already very high. In order not to let the pressure climb any higher, heat supply (instead of combustion) is car-ried out at constant pressure instead of constant volume. The process is thus composed of two isentropes, an isobar, and an isochore, see Fig. 2.17.

qsuppl.pv=const.3pmaxTpmax32qsuppl.4qremov.1TDCBDCv21qremov.4sFig. 2.17: The constant-pressure cycle in the p,v and T,s diagram Again, for the thermal efficiency applies 222 Introduction into the functioning of internal combustion engines

ηth,p=1−

qremovedc(T−T1)

=1−v4 .

qsuppliedqsupplied

As opposed to the constant-volume process, there now appear however three prominent vol-umes. Therefore, a further parameter for the determination of ηth,p is necessary. Pragmati-cally, we select

q*=

qsuppliedcpT1

for this. With this, we first obtain

ηth,p=1−

·cv(T4−T1)1§T4

¨¸=1−−1¸ . cpT1q*κq*¨T©1¹

ꪧq*º·

«¨» . ¸−+11¸κ−1«¨»¹¬©ε¼

And finally after a few conversions

ηth,p

1

=1−

κq*

(2.49)

The thermal efficiency profile of the constant-pressure process in contingency on ε and q*is represented in Fig. 2.18.

0.80.7Thermalefficiencyhthhth,V12hth,p0.60.50.40.30.20.100124681012Compressionratioe14161820q*=4.5719.142Fig. 2.18: Thermal efficiency of the constant-pressure cycle

• The Seiliger cycle

The Seiliger cycle, demonstrated in Fig. 2.19, represents a combination of the constant-volume and the constant-pressure processes.

2.3 Thermodynamics of the internal combustion engine 23

p3241v3*pmaxT213*34sFig. 2.19: The Seiliger cycle in the p,v and T,s diagram

One utilizes this comparative process when, at a given compression ratio, the highest pressure must additionally be limited. The heat supply (instead of combustion) succeeds isochorically and isobarically. With the pressure ratio π=p3p1, we finally obtain the relation

ηth,vp

1

=1−

κq*

κκ−1­ª½1πº§1·°°κ(π−ε)+»¨¸−1¾®«q*−

ε¼©π¹κε°°¯¬¿

(2.50)

for the thermal efficiency, which is graphically represented in Fig. 2.20. From this it becomes

clear that, at a constant given compression ratio ε, it is the constant-volume process, and at a constant given pressure ratio π, it is the constant-pressure process which has the highest efficiency. 0.8Constant-volume0.6hth0.4Constant-pressure0.2p=100p=50p=150p=2000.80.70.6hth0.50.40.30.20.1200150p10050102030e051015e2025300Fig. 2.20: Thermal efficiency of the Seiliger cycle

242 Introduction into the functioning of internal combustion engines

• Comparison of the cycles

p33'21ε= const.qsuppl= const.3332 41BDCvv = const.TDCTε= const.qsuppl=const.33'p = const.132v = const.4additional∆qremovedin the constant-pressure cycleadditional∆removedin the Seiliger cycles21TDCBDCFig. 2.21: Comparison of the closed cycles, ± = constant-volume, ² = constant-pressure, ³ = Seiliger

cycle

For the efficiency of the particular comparative processes, the following contingencies result

1¸ηth,c=f¨¨T¸©3¹

ηth,v=f(ε)

§T·

Carnot,constantvolume,constantpressure,

ηth,p=f(ε,q*)

3

ηth,vp=f¨¨ε,q*,p¸¸

1¹©

§p·

Seiliger.

In Fig. 2.21, the constant-volume, constant-pressure, and the Seiliger processes are repre-sented together in a p,v and T,s diagram. The constant-volume process has the highest, while

2.3 Thermodynamics of the internal combustion engine 25

the constant-pressure process has the lowest efficiency. The efficiency of the Seiliger cycle lies between them. In this comparison, the compression ratio and the supplied heat amount are equal for all three cycles. This shows clearly that, in the case of the Seiliger cycle some, in the case of the constant-pressure process unmistakably more heat must be removed as in the constant-volume process and therefore that the thermal efficiency of these processes is lower.

2.3.3 Open comparative processes

•

The process of the perfect engine

The simple cycles depart partially, yet significantly from the real engine process, such that no detailed statements are possible about the actual engine process. Therefore, for further inves-tigations we also consider open comparative processes, which consider the chemical trans-formation involved in combustion instead of the heat supply and removal of the closed cycles. As opposed to closed cycles, the open comparative cycles allow a charge changing, and they calculate the high pressure process in a stepwise fashion and are thus more or less realisti-cally. However, the charge changing is, as a rule, also not considered more closely. The essen-tial differences to the closed cycles are:

- - -

compression and expansion are either considered isentropically, as previously, or they are described via polytropic state changes.

energy release via combustion is calculated gradually, if also with certain idealiza-tions with reference to the combustion itself.

an approach to the consideration of energy losses resulting from heat transfer is made.

Fig. 2.22 shows an open comparative process in the T,s diagram. Exactly speaking, the fresh load \"appears\" at point 1 and the exhaust gas, which is viewed as mixture burned to an arbi-trary extent, \"disappears\" at point 4.

3T421FreshchargesFig. 2.22: Ideal process for internal combustion engines

Exhaustgasincompletelyburnedtoanarbitraryextend262 Introduction into the functioning of internal combustion engines

The charge changing is not considered here any further. For additional details, refer to Pischinger et al. (2002). •

Heat release through combustion

A distinction is made between incomplete/complete and perfect/imperfect combustion. For air ratios λ≥1, the fuel could in principle burn completely, i.e. the added energy mfuellhv is completely converted into thermal energy

Qmax=Qth=mfuellhv .

For complete combustion of hydrocarbons, the total reaction equations

H2+1O=H2O22

C+O2=CO2.

are valid. Thus, only the products water and carbon dioxide are formed.

In actuality, however, combustion advances maximally until chemical equilibrium for air ratiosλ≥1 as well, thus always incompletely.

For air ratios λ<1, the fuel cannot completely burn as a result of a lack of O2. In the case of such incomplete combustion, the combustion proceeds, in the best case, until chemical equilibrium.

Under all air ratios, combustion can further progress imperfectly, in the case that the oxygen present is not sufficiently optimally distributed (mixture formation), or in the case that single reactions progress slowly and thus that chemical equilibrium is never reached. In the exhaust, we therefore find not only CO2 and H2O, but also carbon monoxide, unburned hydrocar-bons, soot particles, and nitrogen compounds.

1.0ηconv0.90.8aImperfect combustionbcIncomplete combustion0.70.80.91.01.11.21.3λ1.4a:ηconv-loss through lack of oxygenb:ηconv-loss through incomplete combustionc:ηconv-loss through imperfect combustionFig. 2.23: Release of energy and degree of conversion

2.3 Thermodynamics of the internal combustion engine 27

The degree of conversion is defined as

ηconv=1−

Qub

.

mfuellhv

According to Pischinger et al. (19), the total degree of conversion can be written thus:

ηconv,total=ηconv,ch⋅ηconv

withub = unburned and ch = equilibrium.

On the basis of reaction-kinetic estimations, Schmidt et al. (1996) provide the term

­1

¯1.3773λ−0.3773

forfor

ηconv,ch=®

λ≥1λ≤1

for the degree of conversion ηconv,ch. These conditions are clarified visually in Fig. 2.23. •

The real engine process

Proceeding from the perfect engine process, the actual efficiency of the real engine process can be established through gradual abandonment of particular idealizations. Practically, par-ticular losses are thereby considered via corresponding reductions in efficiency, e.g. Volumetric efficiency ∆ηrch: Loss with respect to the entire engine because the actual cylin-der filling at \"intake closes\" is smaller than that of the ideal en-gineCombustion ∆ηic: Heat transfer pwh: Charge changing ∆ηcc: Friction∆ηm:

Loss as a result of incomplete or imperfect combustion Heat losses through heat transfer to the combustion chamber walls

Charge changing losses

Mechanical losses due to engine friction (piston-piston rings-cylinder liner, bearing) and accessory drives (valve train, oil and water pump, injection pump).

Blow-by∆ηbb: Leakage

Particular losses are visually clarified in Fig. 2.24. We will not go into any further detail, one reason being that these simple considerations are of increasingly less importance.

282 Introduction into the functioning of internal combustion engines

0.60.5Efficiency[-]0.40.30.20.10.002∆ηcc∆ηwhSource:Pischinger∆ηrc∆ηicηiLosses:∆ηmMechanical losses∆ηrchReal charge∆ηicImperfectcombustion∆ηrc∆ηwhReal combustionprogressionWall heat transferηe∆ηm∆ηbbBlow-by46810Meaneffectivpressure [bar]12∆ηccCharge changing processηe= ηi− ∆ηmηi= ηV− ∆ηg∆ηg=∆ηrch+∆ηic+∆ηrc+∆ηwh+∆ηbb+∆ηccFig. 2.24: Division of losses in the real internal combustion engine

2.4 Characteristic qualities and characteristic values

Characteristic quantities of internal combustion engines are important with reference to the interpretation and determination of engine measurements, the examination and establishment of the actual performance and the evaluation and comparison of various combustion ma-chines.

Themedium pressure is an important characteristic quantity for the evaluation of performance and of the technological standing of an internal combustion engine. From the definition of piston work

dW=pApdx=pdV

we obtain through integration over a working cycle for the indicated work per cycle

Wi=

³pdV

and with the definition

Wi=imepVd

for the indicated mean effective pressure

imep=

1Vd

³pdV .

(2.51)

For the indicated or internal performance of a multi-cylinder engine follows

Pi=Pi,cylz=znwcimepVd .

2.4 Characteristic qualities and characteristic values 29

With the number of working cycles per time

nwc=i⋅n

with

­0.5i=®

¯1

for4−strokefor2−stroke

we finally obtain for the indicated total performance for

Pi=iznimepVd .

(2.52)(2.53)

Analogous to this, we obtain, with the actual mean effective pressure mep

Pe=iznmepVd

for the actual total performance. The actual performance is the difference from indicated performance and friction performance

Pe=Pi−Pfric ,

(2.)(2.55)

from which follows the relation

fmep=imep−mep

for the friction mean effective pressure.

The internal performance of an engine follows from the so-called indicator diagram, the ac-tual performance follows from

Pe=T2πn

(2.56)

whereby the torque T and the speed n are determined on an engine torque stand.

Theefficiency of a thermal energy conversion machine is quite generally the relation of bene-fit to expenditure. In the case of the internal combustion engine, the benefit the indicated or actual engine performance and the expenditure the energy added with the fuel mass flow 󰀆fuellhv. With this, it follows m

ηi,e=

Pi,e

󰀆fuellhvm

.

(2.57)

The ratio of actual to indicated performance is the mechanical efficiency

ηm=

Pηemep

=e= . ηiPíimep

(2.58)

Thespecific fuel consumption is the fuel consumption related to the engine performance

be=

󰀆fuelmPe

=

1

ηelhv

(2.59)

With a medium value for the lower heating value of gasoline and diesel oil of about

ªkJº

lhv≈42,000«»

¬kg¼

302 Introduction into the functioning of internal combustion engines

one obtains the simple rule-of-thumb between the specific fuel consumption and the actual efficiency

be≈

86ªgº

» . ηe«¬kWh¼

An actual efficiency of ηe=40%, for example, leads to a specific fuel consumption of

be=215gkWh.

We distinguish between upper and lower heating values. In the determination of the upper heating value, the combustion products and cooled back to intake temperature, the water contained in it is condensed out and is thus liquid. As opposed to this, in the determination of the lower heating value, the water is not condensed out and is thus in a vapor state. For inter-nal combustion engines, the lower heating value must be used because of the relatively high exhaust gas temperature.

A so-called mixture heating value is occasionally utilized, under which is meant the added energy flow in reference to the fresh charge. For SI and diesel engines, we retain different expressions for this, because one takes in a gasoline-air mixture and the other pure air.

Hm=Hm=

mfuellhvVmmfuellhvVair

withwith

Vm=Vair=

mair+mfuel

ρm

mair

SIenginedieselengine.

ρair

The ratio of added fresh charge to the theoretically possible charge mass mth is designated as air expenditure,

λair=

mmmm

= . mthVdρth

(2.60)

The theoretical charge density is the density of the intake valve.

As opposed to the air expenditure, the volumetric efficiency designates the ratio of charge mass actually found in the cylinder after charge changing in comparison to the theoretically possible charge mass.

λL=

mcylmth

=

mcylVdρth

. (2.61)

Volumetric efficiency λL is contingent above all on valve overlap in the charge changing-TDC. An optimization of the volumetric efficiency can, with constant control times, only ensue for one speed. With variable valve timing (e.g. by retarding the camshaft), volumetric efficiency can be optimized across the entire speed range. For 4-stroke engines with little valve overlap, λL≈λa is valid.

In addition to the above-cited quantities, a few other characteristic quantities are used. The medium piston speed is a characteristic speed for internal combustion engines,

cm=2sn .

(2.62)

2.5 Engine maps 31

The maximal piston speed is dependant on the eccentric rod ratio and lies in the range cmax=(1.6−1.7)cm.

The compression ratio is the total cylinder volume in reference to the compression volume,

VdVc

ε=1+

(2.63)

The displacement (cubic capacity) is the difference between total volume and compression volume. With the piston path s results for this

Vd=

π4

D2s .

(2.)

As a further characteristic quantity, the bore/stroke ratio

sD

(2.65)

is used.

2.5 Engine maps

2.5.1 Spark ignition engines

In the case of the conventional spark ignition (SI) engine, gasoline is sprayed in the intake port directly in front of the intake valve (multi point injection). Through this, a mixture of air and fuel is taken in and compressed after the closure of the intake valve. Before arrival at the top dead center, the compressed mixture is ignited by means of a sparkplug (spark ignition). Because two phases, the intake and compression strokes, are available for mixture formation, the mixture is nearly homogenous at the end of compression. In conventional engines, the amount of inducted air is regulated by means of a throttle valve positioned in the suction line. At light throttle conditions, this throttle valve is almost closed, while at full loads is it com-pletely open. The amount of inducted air is measured and the fuel injected in proportion to the amount of air, normally such that the medium air ratio λ=1 is maintained (see chapter 4). Because the amount of mixture is regulated in SI engines, we speak of a quantity regulation. So that no autoignition takes place in the compressed mixture, the compression ratio ε has to be restricted.

Fig. 2.25 shows the p,v diagram for a 4-stroke SI engine at partial load (left) and full load (right). The throttle valve, nearly closed at partial load, results in high pressure losses in the suction hose. It thus leads to a \"large charge changing slide\" and finally to inferior efficiency.

322 Introduction into the functioning of internal combustion engines

a)p3b)Expansionandflowlosses2EVO4pExpansionandflowlossesThrottlelossesEVOIVOpupu1'IVOEVCTDCIVC1BDCvIVCEVCTDCBDCvFig. 2.25:p,v diagram for a 4-stroke SI engine under a) full load and b) partial load Fig. 2.26 shows the engine map for a 4-stroke SI engine. The characteristic map is restricted by the idle and limit speeds as well as by the maximal torque line. Because P~T⋅n, the lines of constant performance are hyperbolas in the engine map. The so-called chondiodal curves are lines of constant specific consumption. 14Meaneffectivepressure[bar]12108201000200030004000Enginespeed[rpm]241246260265270300340400[g/kWh]50006000Fig. 2.26: Engine map for a 4-stroke SI engine with lines of constant consumption The behavior and characteristics of the internal combustion engine can be read off from the engine map. From the relations for performance and for torque (load) Pe=inzmepVd=Te2πnwe obtain the contingencies 2.5 Engine maps 33

power:Pe~nmepload:

Te~mep.

The load thereby corresponds to the torque and not to performance! For the ratio of actual to indicated fuel consumption follows

Pe+PfricPfricbePfmep

. =i==1+=1+

biPePePemep

The friction mean effective pressure fmep is nearly proportional to the speed. Therefore at a

constant speed, be must increase with a sinking actual medium pressure.

FlexibilityF of the internal combustion engine is the slope of the torque at the rated load point

§dT·

. F≡−¨¸

©dn¹nrated

(2.66)

As an approach,

F*≡

TmaxnratedTratednTmax

(2.67)

is also often used.

In a flexible engine, the speed lowers under increasing loads, but under constant performance less so as in an inflexible one.

2.5.2 Diesel engines

In the case of the conventional diesel engine, only air is taken in and compressed. The fuel (diesel oil) is injected just before the top dead center into the hot air. Because of the high air ratios, the temperature of the compressed air is clearly higher than the autoignition tempera-ture of the fuel, and after the so-called ignition lag time (see chapter 4) autoignition begins. In contrast to the SI engine, no homogeneous mixture can form in the short time between injec-tion start and autoignition: injection, mixture formation, and combustion therefore proceed in partial simultaneity. The regulation of the diesel engine ensues with the amount of injected fuel, one thus refers to a quality regulation. While in the conventional SI engine the air ratio is alwaysλ=1, it varies in the case of the diesel engine with the load and moves within the region1.1≤λ≤10.

Fig. 2.27 shows the engine map of a 4-stroke diesel engine. We recognize that the speed spread is clearly narrower and that the actual medium pressure is evidently higher than in the case of the SI engine.

342 Introduction into the functioning of internal combustion engines

16141210Specificconsumption[g/kWh]8201000Meaneffectivepressure[bar]15002000250030003500Enginespeed[rpm]40004500Fig. 2.27: Engine map of a 4-stroke diesel engine

With reference to their speeds, internal combustion engines are divided into high-speed en-gines, medium-speed engines, and slow-speed engines, with reference to their construction sizes, however, into vehicle, industrial, and large engines as well. On the other hand, engines built for racing have a special position, particularly because extremely light construction and high performance are principally stressed. With that we have the classification provided in Fig. 2.28.

TypePassenger cars - SIPassenger cars - dieselTrucks - dieselHigh-speed enginesMedium-speed enginesSlow-speed enginesRacing enginesnrpm< 7000< 5000< 3000mepbar8 - 137 - 2215 - 25ηe0.25 - 0.350.30 - 0.400.30 - 0.450.30 - 0.45< 0.5< 0.55- 0.3ε6 - 1216 - 2210 - 2211 - 2011 - 1511 - 157 - 11cmm/s9 - 209 - 169 - 147 - 125 - 105 - 7< 251000 - 250010 - 30150 - 100015 - 259 - 1550 - 15012 - 35Fig. 2.28: Classification of internal combustion engines

2.6 Charging 35

This classification is certainly not compulsory and is not without a certain amount of arbi-trariness, yet it is still practical and comprehensible. We could, in principle, expand it further in consideration of the categories mini-engines (model air planes), small engines (chain saws), and motorcycle engines.

2.6 Charging

Charging was originally considered a performance improvement method. It has, however, been playing an increasingly large role, whereby consumption and emission questions have stepped more into the foreground. For an exhaustive treatment, the reader is referred to Zin-ner (1985) and Pischinger et al. (2002). Jenni (1993) has provided an interesting portrait of the historical development of charging.

2.6.1 Charging methods

We distinguish between external and internal charging, see Fig. 2.29.

ExternalchargingmethodInternalchargingmethodResonance/swingpipechargingMechanicalsuperchargingExhaustturbocharging362 Introduction into the functioning of internal combustion engines

--

mechanical charging,

whereby a supercharger/compressor mounted on the engine is run,

turbocharging (TC),

whereby a charge-air compressor, under the utilization of exhaust gas energy, is run by means of an exhaust turbine, i.e. the reciprocating engine is linked with the flow machine \"turbocharger\" only fluid-mechanically, see Fig. 2.30.

Compressor/Turbinep1nC=nTp5.mairp1cyl.mexhp4cylp6·Highflowrate·Lowpressureratiop2·Lowflowrate·HighpressurerationMEngineFig. 2.30: Reciprocating engine and turbo-machine. Indices: 1: b.c., 2: a.c., 3: b.t., 4: a.t. In the case of turbocharging, we distinguish between: - constant-pressure turbocharging, whereby only the thermal energy of the exhaust gas is utilized; and under this we distinguish between 󰂃󰀃󰂃󰀃single-stageand two-stage charging - and pulse turbocharging, whereby both the thermal and the chemical energy of the exhaust gas is utilized. several equally large turbochargers, or several turbochargers of varying sizes Register charging is a one or two-stage constant-pressure turbocharging method, in which - - are connected one after the other, i.e. with ascending engine load and speed. It is employed in large medium-speed and high-speed high performance diesel engines, two-stage register constant-pressure turbocharging, however, is only used in high-speed high performance die-sel engines. 2.6 Charging 37

We understand under the concept of composite method a combination of varying charging methods in one and the same engine, e.g.:

- - -

mechanical charging for light load operation, TC for medium load operation, and

TC and effective turbine for full load operation.

Fig. 2.31 shows the principle sketch of single- and two-stage turbocharging with intercooling. Under the presumption that the same boost pressure is reached respectively, as a result of intercooling (= low pressure-charge air cooler), the compression work in two-stage charging is smaller than in the single-stage method; correspondingly, the total amount of heat to be removed is also smaller. Under utilization of infinitely many intercoolers, an isothermal com-pression would theoretically be realized.

AWt,AWt,121DELowpressurelevelq21Highpressurelevelq1BCWt,22q22Fig. 2.31: Single- and two-stage exhaust turbocharging 2.6.2 Supercharging In supercharging, the engine and compressor speeds are rigidly coupled. The actual perform-ance of the engine while running a mechanically driven supercharger comes out to Pe=Piηm−Pc1ηg11(2.68)with a transmission efficiency ηg. For the compressor performance, we obtain 󰀆c∆his,cPc=mηis,cηm,c . (2.69)382 Introduction into the functioning of internal combustion engines

Should the flow velocities in front of and behind the compressor be about the same, it then follows from the 1st law of thermodynamics for the isentropic downhill grade

∆his,c

κ−1ºª

«§pa.c.·κ»

¸=cpTb.c.«¨1−»¨¸

«©pb.c.¹»¬¼

and for the isentropic efficiency

ηis,c=

∆his,c∆hc

=

Ta.c.s−Tb.c.

.

Ta.c.−Tb.c.

With these relations, we obtain for the temperature ratio at the compressor

Ta.c.1

=1+Tb.c.ηis,c

κ−1ªº

§·κ«pa.c.»

¸1−«¨» . ¨p¸

«©b.c.¹»¬¼

(2.70)

2.6.3 Constant-pressure turbocharging

In turbocharging, the reciprocating engine (displacement machine) is only connected to the

TC (flow machine) fluid-mechanically, and therefore the speed of the TC is perfectly inde-pendent of the engine speed. In the case of the stationarily running TC, the performance pro-vided by the turbine must be equal to that received by the compressor. For the performance balance at the TC is thus valid, see also Fig. 2.32,

Pc=Pt

(2.71)

1

with the performance of the compressor from (2.69)

󰀆c∆his,cPc=m

ηis,cηm,c

(2.72)

and the performance of the turbine

󰀆t∆his,tηis,tηm,t , Pt=m

if the friction performance of the wheel assembly in the mechanical efficiency of the turbine

and the compressor are taken into account.

In Fig. 2.33, the conditions before and after the compressor and turbine in the h,s diagram are illustrated.

2.6 Charging pC=p2/p1pT=p5/p6p2p5nCp1p6Fig. 2.32: Performance balance of a turbocharger, Indices: 1: b.c., 2: a.c., 3: b.t., 4: a.t. h22isCompressorDhDhishDhisisC=Dh1shTurbine5hDhisT=DhisDhisDh6is6sFig. 2.33:h,sdiagram for compressor and turbine By equating the compressor and turbine performances, it follows from (2.71) m󰀆tm󰀆η∆his,ccis,tηis,cη󰀉m󰀌,t󰀊η󰀌m,c= . 󰀉󰀌󰀌󰀌󰀊󰀌η󰀋∆his,t󰀌n,TC󰀉󰀌󰀌󰀌󰀌󰀊η󰀌󰀋󰀌TC󰀌󰀌󰀌󰀋η*TC39

(2.73)402 Introduction into the functioning of internal combustion engines

With the isentropic downhill grade for the compressor

§Ta.c.·

¸=cpc∆his,c=cpcTb.c.¨−1¨T¸

©b.c.¹

κc−1ªº

§·κ«pa.c.»

¸c−1»Tb.c.«¨¨¸

«©pb.c.¹»¬¼

(2.74)

and for the turbine

§Ta.t.·

¸=cpt∆his,t=cptTb.t.¨−1¨¸Tb.t.¹©

κt−1ºª

«§pa.t.·κt»

¸Tb.t.«1−¨»¨p¸bt..©¹«»

¬¼

(2.75)

one finally obtains the freewheel condition or also the so-called 1st fundamental equation of

turbocharging

−1¨p¸Tbc..¹*b.t.

=©ηTC .

κt−1cpcTb.c.

§pa.t.·κt

¸1−¨¨p¸©b.t.¹cpt

κc−1

§pa.c.·κc¨¸

(2.76)

For the pressure ratio at the compressor, it follows from this

§pb.t.*Ta.t.·pa.c.

¸=f¨¨p,ηTCT¸ . pb.c.b.c.¹©a.t.

(2.77)

Because the composition of the fresh air flowing through the compressor and the exhaust gas

flowing through the turbine is different, a differentiation must be made between material valuescp and κ for both of these gas flows, which should be expressed by the additional

󰀆c and m󰀆t and the specific heat capacities cpcindicesc and t. Since both the mass flows m

andcpt differ only slightly, an approach to considering these differences can be made in the modified efficiency η*TC.

The mass flow through the turbine, the so-called turbine absorption capacity, can be roughly determined with the help of the relation for flow through restriction (throttle). Thus, with the isentropic equivalent cross section of the turbine, we obtain

󰀆t=Ais,tρisvism

with

1

ρis=

and

pb.t.

RtTb.t.

§pa.t.·κt¨¸¨p¸©b.t.¹

b.t..mEa.t.2.6 Charging 41

vis=

κt−1ºª

«§pa.t.·κt»2κt

¸RtTb.t.«1−¨» . ¨p¸κt−1bt..©¹«»

¬¼

󰀆t and conversion finally provides the expression Substitutingρis and vis in the relation for m

󰀆tm

Tb.t.pb.t.

=Ais,t

2κt+1ºª

«§pa.t.·κt§pa.t.·κt»2κt

¸−¨¸» . «¨¸¨p¸Rt(κt−1)«¨p©b.t.¹©b.t.¹»¬¼

(2.78)

This is the so-called 2nd fundamental equation of turbocharging

󰀆tm

Tb.t.pb.t.

§pa.t.·¸ . =Ais,tf¨,κt¨p¸

©b.t.¹

(2.79)

With the help of this relation, reference or dimensionless quantities can be produced, with

which we will deal further in chapter 7.

The isentropic turbine equivalent cross section is approximately constant and contingent only on the geometric turbine cross-section; dependence upon the speed of the turbocharger and the pressure ratio pb.t.pa.t. is thereby ignored.

We immediately see that the chosen turbine cross-section and thus the chosen turbine is only optimal for a certain operation point. For lower engine speeds as the selected one, a turbine with a smaller cross-section would be necessary, for higher speeds a larger turbine. At lower speeds, an excessively large turbine leads to a torque/performance drop. A modulation valid for a larger speed range can be achieved through

- - -

charge-air release (boost pressure restriction) exhaust gas release (wastegate)

a turbine with an adjustable geometry (VTG charger), i.e. with a variable turbine cross-section

- register charging.

2.6.4 Pulse turbocharging

In pulse turbocharging, we keep the volume of the exhaust pipe very small. In this way, a very fast filling of the pipe is achieved and a large part of the kinetic energy of the exhaust gas is thereby retained. As opposed to this, in the case of constant-pressure turbocharging, the con-version of kinetic energy into pressure energy is fraught with losses. In Fig. 2.34, pressure and pulse turbocharging are presented in an h,s diagram. The principle differences are clearly recognizable. The enthalpy downhill grade at the turbine is evidently larger in the case of pulse turbocharging. Yet the turbine is unsteadily admitted.

Because the pressure in the exhaust pipe is no longer constant, in the case of multicylinder engines, one can only amalgamate those cylinders in which the pressure impacts in the ex-haust system do not reciprocally disturb the charge changing. Therefore, the ignition distance

422 Introduction into the functioning of internal combustion engines

from the amalgamated cylinders should be larger than the opening duration of an exhaust valve, from which, according to experience, an ignition distance of at least 240° CA results for a disturbance-free region.

hh°4hh°44p666s6Pressurechargingp°44p45p5p°4p4p°5p5p6DhTDhisT66sPressuretube45DhTDhisT4hushu5Pulsetubes6PulsechargingFig. 2.34: Pressure and pulse turbocharging in the h,s diagram Fig. 2.35 shows the diagram of a pulse turbocharged 6-cylinder 4-stroke diesel engine, whereby the exhaust pipes of cylinder 1, 2, and 3 as well as those of 4, 5, and 6 are respec-tively amalgamated. Both of these multiple vents are lead separately to the turbine. The tur-bine is equipped with a so-called twin spiral casing. Separate exhaust pipe from the engineOil supplyto the turbine of the turbocharger.from the engineTurbineCompressorAmalgamation of the cylindersaccording to the firing order of the engine.Exhaust gassesIntake air throughExhaust gasto the turbinethe air filterexitof the turbochargerto the turbochargerOil drainto theengineCompressed airto the engineFig. 2.35: Pulse turbocharged 6-cylinder, 4-stroke engine 2.6 Charging 43

In conclusion, Fig. 2.36 shows the engine map realizable with varying charging methods for a fast-running high performance diesel engine. With two-stage constant-pressure turbocharging in register mode, we see a very broad map, with single-stage pulse turbocharging, on the other hand, a relatively constant torque at high performance. 可调涡轮?Pe, max100%80Two-stage pressurecharging withturbo shiftingTorquepelle40rgrapSingle-stageshock chargingPro20hP~n60P~ n2ExcessspeedTwo-stage pressurecharging withoutturbo shifting3Drag torque020Idle4060Engine speed80%100Fig. 2.36: Comparison of different charging methods