DieselDieselEngine Engine -CyclesCycles€¦ · Qianfan Xin, Diesel engine system design, Woodhead Publishing Limited, 2011 Diesel engineclassification. In external combustion, the

DieselDiesel Engine Engine -- CyclesCycles

A quantity of heat is conveniently measured by applying it to raise thetemperature of a known quantity of pure water.The unit of heat is defined as that quantity of heat required to raise thetemperature of unit weight of water through one degree, this quantitydepending, of course, on the particular unit of weight and the temperaturescale employed.The Continental European and scientific temperature scale has been theCentigrade scale, now called Celsius because of possible confusion with theFrench meaning of the word centigrade – one ten thousandth of a right angle.The interval between the temperatures of melting ice and boiling water (atnormal pressure) is divided into one hundred, though the unsatisfactoryFahrenheit scale, which divides the foregoing interval into 180 divisions, hasbeen the commercial standard in Britain and the USA.

Heat and work

T.K. GARRETT, K. NEWTON, “The Motor Vehicle”, Thirteenth Edition, 2001

WorkIf work is done by rotating a shaft, the quantity of work is the product of thetorque or turning moment applied to the shaft in newton metres, multiplied bythe angle turned through measured in radians. One revolution equals 2πradians.

Joule’s equivalentDr Joule was the first to show, in the middle of the last century, that heat andwork were mutually convertible one to the other, being, in fact, different formsof energy, and that when a definite quantity of work is expanded wholly inproducing heat by friction or similar means, a definite quantity of heat isproduced. His experiments, confirmed and corrected by others, showed that778 foot-pounds produce one Btu, or 1400 foot-pounds produce one CHU.This figure is called the mechanical equivalent of heat, though it wouldperhaps be better to speak of the thermal equivalent of work. For though thesame equiválent or rate of exchange holds for conversion in either direction,while it is comparatively simple to convert to heat by friction the whole of aquantity of work supplied, it is not possible, in a heat engine, to convert tomechanical work more than a comparatively small percentage of the totalheat supplied. There are definite physical laws which limit this percentage –or thermal efficiency as it is called – to about 50% or less in the best heatengines that it is practicable to construct.

Heat and work


Thermal efficiencyThe thermal efficiency is governed chiefly by therange of temperature through which the workingfluid, be it gas or steam, passes on its waythrough the engine.This range of temperature is greater in internalcombustion engines than in steam engines,hence the former are inherently capable ofhigher thermal efficiencies, that is to say, thryare capable of converting into work a higherpercentage of the total heat of the fuel withwhich they are supplied than the latter. Even so,the physical limitations are such that thethermal efficiency of a good petrol engine is notmore than about 28%. The remaining heatsupplied, which is not converted into work, islost in the exhaust gases andcooling water, and in radiation.

Thermal efficiency


When unit weight of any fuel is completely burnt with oxygen (pure or diluted withnitrogen as in the air), a certain definite quantity of heat is liberated, depending on thechemical composition, that is, on the quantities of the fundamental fuels, carbon andhydrogen, which one pound of the fuel contains.To determine how much potential heat energy is being supplied to an engine in a giventime, it is necessary to know the weight of fuel supplied and its calorific value, which isthe total quantity of heat liberated, when unit weight of the fuel is completely burnt.The calorific values of carbon and hydrogen have been experimentally determinedwith considerable accuracy, and are usually given as —Carbon 33 000 kJ/kg, or 14200 Btu/lbHydrogen 144300 kJ/kg, or 62100 Btu/lbThe calorific value of any fuel, consisting, as all important fuels do, of a knownproportion of carbon, hydrogen and incombustible impurities or diluents, may beestimated approximately on the assumption that it consists simply of a mixture ofcarbon, hydrogen and incombustible matter, but the state of chemical combination inthe actual fuel leads to error by this method, and the only accurate and satisfactorymeans of determination is experimentally by the use of a suitable calorimeter.

Calorific value


Power is the rate at which work is done, 1 hp being defined (by James Watt) as arate of working of 33000 ft lb per minute, or 550 per second. (1 hp = 745.7 W).

Power


Average petrol consists approximately of 85%carbon and 15% hydrogenby weight, the lighter fractions containing ahigher percentage of hydrogen than the heavier.Refined petrol contains no measurableimpurities or diluents.Its gross calorific value is about 46000 kJ/kg, or19 800 Btu/lb.Liquid fuels are usually measured by volume,and therefore it is necessary to know the densitybefore the potential heat supplied in any givencase can be determined, for example—A sample of petrol has a calorific value of 46000kJ/kg; its specificgravity is 0.72. How much potential heat energyis contained in 8 litres?(1cm3 of water weighs 1 g.)Weight of 8 litres = 8 × 0.72 = 5.76 kg.Thus, the potential heat in 8 litres = 5.76 ×46000 = 264 900 kJ.

Calorific value


The internal-combustion (IC) engine isthe most frequently employed power sourcefor motor vehicles. Internal-combustionengines generate power by convertingchemical energy bound in the fuel into heat,and the heat thus produced into mechanicalwork. The conversion of chemical energyinto heat is accomplished throughcombustion, while the subsequentconversion of this thermal energy intomechanical work is performed by allowingthe heat energy to increase the pressurewithin a medium which then performs workas it expands.Liquids, which supply an increase inworking pressure via a change of phase(vaporization), or gases, whose workingpressure can be increased throughcompression, are used as working media.

Operating Concepts and Classifications

The fuels – largely hydrocarbons – requireoxygen in order to burn; the required oxygen isusually supplied as a constituent of the intake air. Iffuel combustion occurs in the cylinder itself, theprocess is called internal combustion. Here thecombustion gas itself is used as the workingmedium.If combustion takes place outside the cylinder, theprocess is called external combustion.Continuous mechanical work is possible only in acyclic process (piston engine) or a continuousprocess (gas turbine) of heat absorption, expansion(production of work) and return of the workingmedium to its initial condition (combustion cycle).If the working medium is altered as it absorbs heat,e.g. when a portion of its constituents serve as anoxidant, restoration of its initial condition is possibleonly through replacement. This is called an opencycle, and is characterized by cyclic gas exchange(expulsion of the combustion gases and inductionof the fresh charge). Internal combustion thereforealways requires an open cycle.


Table 1. Classification of the internal-combustion engine

Qianfan Xin, Diesel engine system design, Woodhead Publishing Limited, 2011

Diesel engine classification

In external combustion, the actualworking medium remains chemicallyunchanged, and can thus be returned toits initial condition by suitable measures(cooling, condensation). This enablesthe use of a closed process.In addition to the main processcharacteristics (open/closed) and thetype of combustion (cyclic/continuous),the various combustion processes forinternal combustion engines can also bedefined according to their air-fuelmixture formation and ignitionarrangements.In external air-fuel mixture formation,the mixture is formed outside thecombustion chamber. In this type ofmixture formation a largely homogenousair-fuel mixture is present whencombustion is initiated, so it is alsoreferred to as homogenous mixtureformation.


In internal air-fuel mixtureformation the fuel is introduceddirectly into the combustionchamber. The later internalcombustion occurs, the moreheterogeneous the air-fuelmixture will be at the timecombustion is initiated. Internalmixture formation is thereforealso called heterogeneousmixture formation. Externalignition designs rely on anelectric spark or a glow plugto initiate combustion. Inautoignition, the mixture ignitesas it warms to or beyond itsignition temperature duringcompression, or when fuel isinjected into air whose boundaryconditions permit evaporationand ignition.


glow plug

The p-V diagramA basic precondition for continuous conversionof thermal energy into kinetic energy is amodification in the condition of the workingmedium; it is also desirable that as much of theworking medium as possible be returned to itsinitial condition.

Cycles

A thermodynamic cycle illustrated using the p-V diagram

For technical applications the focus can rest onchanges in pressure and the correspondingvolumetric variations which can be plotted on apressure vs. volume work diagram, or p-Vdiagram for short.As the figure shows, the addition of heat andthe change in condition of the working mediumthat accompany the progress of the process inthe 12 phase must consume less energythan that required for the 21 phase. Once thiscondition is satisfied the result is an areacorresponding to the process work potential:

The T-S diagramThe temperature entropy, or T-Sdiagram, is used to provide asimilar graphic representation ofthe bidirectional thermal energytransfers in this cyclic process.

Cycles

A thermodynamic cycle illustrated using the T-S- or H-S diagram.

Cycles

The Carnot Cycle

The Carnot cycle in the p-V and T-

S diagrams

The Carnot Cycle

The Carnot cycle in the p-V and T-S

diagrams

Theoretical treatment today involves the following ideal combustion cycles:the constant-volume cycle for all piston engines with periodic combustion andgeneration of work, and the constant-pressure cycle for all turbine engines with continuouscombustion and generation of work.Both cycles will be dealt with in more detail in the discussion of the correspondingmachines.1) Isothermal change in condition: temperature does not change.2) Isentropic change in condition: adiabatic (heat is neither added nor dissipated) andfrictionless (reversible).

Reciprocating-piston engines with internal combustion

Operating conceptAll reciprocating-piston engines operateby compressing air or an air-fuel mixturein the working cylinder prior to ignitingthe mixture, or by injecting fuel into thehot compressed air to initiatecombustion. The crankshaft assemblyconverts the work generated in thisprocess into torque available at the endof the crankshaft.

The engine power cycle1 In the p-V diagram,2 in the p-t and p-α diagrams.


The p-V diagram reflects the actual power-generation process in the engine as a function ofpiston travel. It shows the mean effectivepressures pmi within the cylinder during acomplete working cycle. Easier to produce areother diagrams such as the pressure vs. time (p-t-) and the pressure vs. crankshaft angle (p-α-)diagrams. The surfaces defined in these twodiagrams do not directly indicate the amount ofwork generated, but they do provide a clearpicture of essential data such as firing point andpeak combustion pressure. The product of themean effective pressure in the cylinder and thedisplacement yields the piston work, and thenumber of working cycles per unit of timeindicates the piston power or the internal power(power index) for the engine. Here it will be notedthat the power generated by a reciprocatingpiston internal-combustion engine increases asmin^–1 rises


Ideal combustion cycle for piston engineswith internal combustionFor reciprocating-piston engines with internalcombustion, the ideal thermodynamiccombustion process is the "constant-volumeprocess", consisting of isentropiccompression, isochoric) heat supply,isentropic expansion and isochoric reversionof the ideal working gas to its initial condition.This cycle is only possible if thefollowing conditions are met:• No heat or gas losses, no residual gas,• Ideal gas with constant specific heats cp,

cv and χ = cp/cv = 1,4;• Infinitely rapid heat supply and discharge,• No flow losses.

Note: Isochoric change in condition: volumedoes not change.


Ideal constant-volume combustion cycle as shown in the p-V and T-S diagrams.

Because the crankshaft assembly restricts expansion to finite levels, the 4–5–1surface in the diagrams is not directly available for use. Section 4–5'–1, lyingabove the atmospheric pressure line, becomes available when an exhaust-gasturbine is connected downstream.

The spark-ignition engine (or Sl engine) isa piston engine with external or internal air-fuel mixture formation. External mixtureformation generally produces homogenousmixtures, whereas an internally formedmixture is largely heterogeneous at theinstant of ignition. The time of mixtureformation is a major factor influencing thedegree of homogenization achievable byinternal mixture formation.In both cases, the mixture is compressed toapproximately 20...30 bar (ε = 8...12) on thecompression stroke, to generate a finalcompression temperature of 400...500 °C.This is still below the auto-ignition thresholdof the mixture, which then has to be ignitedby a spark shortly before the piston reachesTDC.

The spark-ignition (Otto) engine

Ideal Otto CycleIn an Otto cycle, a fuel–air mixture is introduced duringthe changing of the cylinder, and near the top deadcenter (TDC), the fuel–air mixture is ignited electricallywith the help of a spark-ignition system. Ideally, thecombustion is instantaneous and can be considered aconstant-volume process. It is therefore also called aspark-ignition (SI) engine. Hence, the thermodynamiccycle consists of a four-part process, shownschematically in Fig. 1, as follows: Air is brought into thecylinder at state 1 (intake stroke); it is compressedisentropically (with constant entropy) to state 2(compression stroke); combustion takes place at aconstant volume to reach state 3; it is expandedisentropically to state 4 ( power stroke); the exhaust gasis expelled (exhaust stroke), and simultaneously a freshchange in the fuel–air mixture is introduced at a constantvolume (isochore). This is when all four processes arecompleted within one cycle in two half-cycles (two-strokeengines), as in the present ideal case, or in four half-cycles; the last two are separate processes for thesuction of air and expelling the hot gas (four-strokeengine).In addition to the compression ratio let the otheroperational parameter be, which depends on the fuel–airratio. Now, for the two isentropic processes,

Fig. 1: Thermodynamicprocess in Otto cycle of a

piston engine

Ideal Otto Cycle

Ideal Otto Cycle

Ideal Otto Cycle

Ideal Otto Cycle21 QQLç

1

2

1

21

1

1QQ

QQQ

QLç

t ).(. 231 TTcmQ v

).(.).(.1

23

14

TTcmTTcm

v

vt

).(. 142 TTcmQ v

)1(

)1(1

3

23

4

14

TTT

TTT

t

Q1=Sisteme giren ısı

Q2=Sistemden çıkan ısı

3

4

33

4.4

3

4

33

44

. TT

VPVP

mRTmRT

VPVP 2314 , VVVV

3

4

2

2

2

1

VV

VVV

VV H kk VPVP 4433 ..

Ideal Otto Cyclekkk

VV

VV

PP

1

1

2

4

3

3

4

.1.1

3

4

3

4kk

TT

VV

13

4 1 kT

T

2

1

2

.1

2

1

22

1.1

2

1

22

11 ... T

TPP

TT

VPVP

mRTmRT

VPVP

kkkk

VV

PPVPVP

1,..1

2

2

12211

.1.1

2

1

2

1kk

TT

VV

12

1 1 kT

T

Ideal Otto Cycle

3

2

4

1

3

4

2

1 ,TT

TT

TT

TT

13

4

3

23

3

24

3

23

4

14 111

)1(

)1(1

)1(

)1(1

kt TT

TTT

TTT

TTT

TTT

Ideal Otto Cycle - Mean Effective Pressure

21 QQLç 1

1

.; QLQL

tçç

t

).(. 231 TTcmQ v

1...2

32 TTTcmL vtç

kcc

v

p vp ckc .1

. kRcRckcRcc vvvvp

112 . kTT

nıpatlamaoraTT

PPVV

mRTmRT

VPVP :

. 2

3

2

.323

2

3

22

33

Ideal Otto Cycle - Mean Effective Pressure

1...1

1.1..1

.. 1111

1 VPk

TkRmL k

tk

tç

)1()11()1( 111

2121

VVVVVVVVH

1.

1.

1.

1.

1...1

1.1

1

111

k

t

kt

H

çmi k

P

V

VPk

VL

P

A diesel engine is a reciprocating-pistonengine with internal (and thusheterogeneous) mixture formation and auto-ignition. During the compression strokeintake air is compressed to 30...55 bar innaturally aspirated engines or 80...110 bar insupercharged engines, so that itstemperature increases to 700...900 °C. Thistemperature is sufficient to induce auto-ignition in the fuel injected into the cylindersshortly before the end of the compressionstroke, as the piston approaches TDC. Inheterogeneous processes the mixtureformation is decisive in determining thequality of the combustion which then follows,and the efficiency with which the inductedcombustion air is utilized, and thus indefining the available mean effectivepressure levels.

The Diesel Engine

Idealized Diesel Cycle

In a diesel cycle, fuel isintroduced at a controlledrate to produce a constantpressure combustion, whilethe piston is withdrawn to acylinder volume V3, andunder ideal conditions,further expansion takesplace isentropically. Theprocess is shownschematically in Fig. 1.

Schematic sketch of a compressionignition engine

Compression Ignition EngineThe diesel internal combustion enginediffers from the gasoline powered Ottocycle by using a higher compression ofthe fuel to ignite the fuel rather than usinga spark plug ("compression ignition" ratherthan "spark ignition").In the diesel engine, air is compressedadiabatically with a compression ratiotypically between 15 and 20. Thiscompression raises the temperature to theignition temperature of the fuel mixturewhich is formed by injecting fuel once theair is compressed.The ideal air-standard cycle is modeled asa reversible adiabatic compressionfollowed by a constant pressurecombustion process, then an adiabaticexpansion as a power stroke and anisovolumetric exhaust. A new air charge istaken in at the end of the exhaust, asindicated by the processes a-e-a on thediagram

IdiealIdieal AirAir Standard Standard DieselDiesel Engine CycleEngine CycleSince the compression and power strokes of thisidealized cycle are adiabatic, the efficiency can becalculated from the constant pressure and constantvolume processes. The input and output energiesand the efficiency can be calculated from thetemperatures and specific heats:

It is convenient to express this efficiency in terms of the compression ratio rC = V1/V2 and the expansion

ratio rE = V1/V3. The efficiency can be written

and this can be rearranged to the form

Diesel Engine – Thermodynamic Ideal Cycle Analysis

21 QQLç

1

2

1

21

1

1QQ

QQQ

QLç

t

).(. 231 TTcmQ p ).(. 142 TTcmQ v

oranıgenişönTT

VVPP

mRTmRT

VPVP

g ..:. 2

3

2

.323

2

3

22

33

kg

k

k

k

k

k

VV

PPPPVV

VPVP

VPVP

2

3

1

42314

22

33

11

44

1

2

4

12

4

1

4

1

.414

1

4

11

44 ..

k

kTT

TT

TT

PPVV

mRTmRT

VPVP


12

41

2

4

1

4 . k

kgkk

g TT

TT

PP

)1(

)(.11

)1(

)(.111

).(.).(.1

2

3

2

1

2

4

2

32

2

1

2

42

23

14

TTTT

TT

kTTT

TT

TTT

kkcc

TTcmTTcm

tp

v

p

vt

)1.()1(

.11)1(

)1(.11 1

11

gk

kg

g

kk

kg

t kk


21 QQLç 11

.; QLQL

tçç

t

).(. 231 TTcmQ v

1...2

32 TTTcmL ptç

kcc

v

p vp ckc .

1.

kRcRckcRcc vvvvp

112 . kTT

1...1

.1..1

...1.1

... 111

11

2

32

gk

tgk

ttç VPkkT

kRkm

TTT

kRkmL

1.

1..

1.

1.

1...1

.1

1

111

g

k

t

gk

t

H

çmi

Pkk

V

VPkk

VL

P

Diesel Engine – Mean Effective Pressure

Summary

Air standard diesel engine cycle

Work per minute, power and horsepowerWork per minute, power and horsepower


Work per minute, power and horsepowerWork per minute, power and horsepower


Incidentally, since 1 hp is defined as the equivalent of 550 ft lbf of work per second,it can be shown that the formula for horsepower is precisely the same as that forthe power output in watts, except that p, D and L are in units of 1bf/in2, in and ft,and the bottom line of the fraction is multiplied by 550.Following the formation of the European common market, manufacturers tended tostandardise on the DIN (Deutsche Industrie Norm 70 020) horsepower, whichcame to be recognised as an SI unit. In 1995, however, the ISO (InternationalStandards Organisation) decreed that horsepower must be determined by the ISO1585 standard test method. This standard calls for correction factors differing fromthose of the DIN as follows: 25°C instead of 20°C and 99 kPa instead of 1013 bar,respectively, for atmospheric temperature and pressure, and these make itnumerically 3% lower than the DIN rating. The French CV (chevaux) and theGerman PS (pferdestarke), both meaning ‘horse power’, must be replaced by theSI unit, the kilowatt, 1 kW being 1.36 PS.

Indicated and brake powerIndicated and brake power


The power obtained from the indicatordiagram (that is, (ipo or ihp), and is thepower developed inside the enginecylinder by the combustion of the charge.The useful power developed at theengine shaft or clutch is less than this bythe amount of power expended inovercoming the frictional resistance of theengine itself. This useful power is knownas the brake power output or brakehorsepower (bpo or bhp) because it canbe absorbed and measured on the testbench by means of a friction or fan brake.(For further information on engine testingthe reader is referred to The Testing ofInternal Combustion Engines by Youngand Pryer, EUP.)

Mechanical efficiency


The ratio of the brake horsepower to the indicated horsepower is known as themechanical efficiency.Thus—

Characteristic speed power curves


If the mean effective pressure (mep) and the mechanical efficiency of anengine remained constant as the speed increased, then both the indicated andbrake horsepower would increase in direct proportion to the speed, and thecharacteristic curves of the engine would be of the simple form shown in Fig. 1,in which the line marked ‘bmep’ is the product of indicated mean effectivepressure (imep) and mechanical efficiency, and is known as brake meaneffective pressure (bmep). Theoretically there would be no limit to thehorsepower obtainable from the engine, as any required figure could beobtained by a proportional increase in speed. It is, of course, hardly necessaryto point out that in practice a limit is imposed by the high stresses and bearingloads set up by the inertia of the reciprocating parts, which would ultimatelylead to fracture or bearing seizure.Apart from this question of mechanical failure, there are reasons which causethe characteristic curves to vary from the simple straight lines of Fig. 1, andwhich result in a point of maximum brake horsepower being reached at acertain speed which depends on the individual characteristics of the Engine.



Fig. 1



Characteristic curves of an early four-cylinder engine of 76.2 mm bore and 120.65 mmstroke are given in Fig. 2. The straight radial lines tangential to the actual power curvescorrespond to the power lines in Fig. 1, but the indicated and brake mean pressures do not,as was previously assumed, remain constant as the speed increases.On examining these curves it will be seen first of all that the mep is not constant. It shouldbe noted that full throttle conditions are assumed – that is, the state of affairs for maximumpower at any given speed.At low speeds the imep is less than its maximum value owing partly to carburation effects,and partly to the valve timing being designed for a moderately high speed; it reaches itsmaximum value at about 1800 rev/min, and thereafter decreases more and more rapidly asthe speed rises. This falling off at high speeds is due almost entirely to the lower volumetricefficiency, or less complete filling of the cylinder consequent on the greater drop of pressureabsorbed in forcing the gases at high speeds through the induction passages and valveports.When the mep falls at the same rate as the speed rises, the horsepower remains constant,and when the mep falls still more rapidly the horsepower will actually decrease as thespeed rises. This falling off is even more marked when the bmep is considered, for themechanical efficiency decreases with increase of speed, owing to the greater frictionlosses. The net result is that the bhp curve departs from the ideal straight line more rapidlythan does the ihp curve. The bmep peaks at about 1400 rev/min, the indicated power at3200 and the brake power at 3000 rev/min, where 33.5 kW is developed. Calculations ofbmep P from torque, and vice versa are made using the following formula—



Fig. 2 Power curves of typical early side-valve engine, 3-in bore and 4 3 in stroke 4 (76.2 and

120.65 mm)



This applies to a two-stroke engine, which has one power stroke per revolution.Because a four-stroke engine has only one power stroke every two revolutions, we musthalve result, so we have—

Torque curves


If a suitable scale is applied, the bmep curve becomes a ‘torque’ curve for the engine, thatis, it represents the value, at different speeds, of the mean torque developed at the clutchunder full throttle conditions – for there is a direct connection between the bmep and thetorque, which depends only on the number and dimensions of the cylinders, that is, on thetotal swept volume of the engine. This relationship is arrived at as follows— If there are ncylinders, the total work done in the cylinders per revolutionis—

Torque curves


Torque curves


It is more usual to calculate the bmep (which gives a readier means of Power – kWcomparison between different engines) from the measured value of the torque obtainedfrom a bench test. Indicated mean pressure and mechanical efficiency are difficult tomeasure, and are ascertained when necessary by laboratory researches. Mean torque, onthe other hand, can be measured accurately and easily by means of the variouscommercial dynamometers available. The necessary equipment and procedure are ingeneral use for routine commercial tests. It is then a simple matter to calculate from themeasured torque the corresponding brake mean pressure or bmep—

The usual form in which these power or performance curves are supplied by the makers isillustrated in Figs 3 and 4, which show torque, power, and brake specific fuel consumptioncurves for two Ford engines, the former for a petrol unit and the latter a diesel engine. Inboth instances, the tests were carried out in accordance with the DIN Standard 70020,which is obtainable in English from Beuth-Vertrieb GmbH, Berlin 30. The petrol unit is anoverhead camshaft twin carburettor four-cylinder in-line engine with a bore and stroke of90.8 by 86.95 mm, giving a displacement of 1.993 litres. Its compression ratio is 9.2 to 1.The diesel unit is a six-cylinder in-line engine with pushrod-actuated valve gear and havinga bore and stroke of 104.8 mm by 115 mm respectively, giving a displacement of 5948cm3. It has a compression ratio of 16.5 :1.

Torque curves


Fig. 3 Typical performance curves for an overhead camshaft, spark-ignition engine. High speeds areobtainable with the ohc layout

Torque curves


Fig. 4 Performance curves of a diesel engine. The fact that torque increases as speed falls off from themaximum obviates the need for excessive gear-changing

Brake specific fuel consumption


When the simple term specific fuel consumption is used it normally refers to brake specificfuel consumption (bsfc), which is the fuel consumption per unit of brake horsepower. InFigs 3 and 4 the specific fuel consumption is given in terms of weight. This is moresatisfactory than quoting in terms of volume, since the calorific values of fuels per unit ofvolume differ more widely than those per unit of weight. It can be seen that the specificfuel consumption of the diesel engine is approximately 80% that of the petrol engine,primarily due to its higher compression ratio. Costs of operation, though, depend not onlyon specific fuel consumption but also on rates of taxation of fuel. The curves show that thelowest specific fuel consumption of the diesel engine is attained as the fuel : air ratioapproaches the ideal and at a speed at which volumetric efficiency is at the optimum. Inthe case of the petrol engine, however, the fuel : air ratio does not vary much, and thelowest specific fuel consumption is obtained at approximately the speed at whichmaximum torque is developed – optimum volumetric efficiency.The fuel injection rate in the diesel engine is regulated so that the torque curve risesgently as the speed decreases. A point is reached at which the efficiency of combustiondeclines, with rich mixtures indicated by sooty exhaust. This torque characteristic isadopted in order to reduce the need for gear changing in heavy vehicles as they mountsteepening inclines or are baulked by traffic. The heavy mechanical components, includingthe valve gear as well as connecting rod and piston assemblies of the diesel engine, andthe slower combustion process, dictate slower speeds of rotation as compared with thepetrol engine.



In Figs 3 and 4, the curves of specific fuel consumption are those obtained when the engineis run under maximum load over its whole speed range. However, in work such as matchingturbochargers or transmission systems to engines, more information on fuel consumption isneeded, and this is obtained by plotting a series of curves each at a different load, or torque,as shown in Fig. 5. Torque, however, bears a direct relationship tobmep and, since this is a more useful concept by means of which to make comparisonsbetween different engines, points of constant bsfc are usually plotted against engine speedand bmep, the plots vaguely resembling the contour lines on an Ordnance Survey map.The curves in Fig. 5 are those for the Perkins Phaser 180Ti, which is the turbocharged andcharge-cooled version of that diesel engine in its sixcylinder form. Such a plot is sometimesreferred to as a fuel consumption map. Its upper boundary is at the limit of operation abovewhich the engine would run too roughly or stall if more heavily loaded; in other words, it is thecurve of maximum torque – the left-hand boundary is the idling speed, while that on the rightis set by the governor. For a petrol engine, the right-hand boundary is the limit beyond whichthe engine cannot draw in any more mixture to enable it to run faster at that loading.Over much of the speed range, there are two speeds at which an engine will run at a givenfuel consumption and a given torque. A skilful driver of a commercial vehicle powered by thePhaser 180Ti will operate his vehicle so far as possible over the speed range from about1200 to 2000 rev/min, staying most of the time between 1400 and 1800 rev/min, to keep hisfuel consumption as low as possible. The transmission designer will provide him with gearratios that will enable him to do so, at least for cruising and preferably over a wider range ofconditions, including up- and downhill and at different laden weights.



Fig. 5 Fuel consumption map for the Perkins Phaser 180Ti turbocharged and aftercooleddiesel engine, developing 134 kW at 2600 rev/min

Table: Graphic representations and definitions of the individual and overall efficiencies of the reciprocating-piston engine

Karma Çevrim (Seilinger)

OKUMA PARÇASI:• Sıkıştırma (1-2)• Bu safhada, piston alt ölü noktadan üst ölü noktaya doğru hareket eder. Busırada emme ve egzoz valfleri kapalıdır, dolayısıyla içerdeki hava sıkışır vebasıncı grafikte görüldüğü gibi artar.• Sabit Hacimde Yanma (2-3)• Piston üst ölü noktaya ulaştığı sırada silindire enjektör tarafından yakıtpüskürtülmeye başlar. Sıkışarak ısınmış havayla karşılaşan yakıt yanmayabaşlar, bunun sonucunda basınç P2'den P3 değerine sıçrama yapar. Sistemeısı girişinin olduğu ilk safha bu safhadır.• Sabit Basınçta Yanma (3-4)• Bu safhada piston aşağı doğru hareketine başlar fakat yanma devamettiğinden basınç düşmez. Bu durum 4 nolu noktaya kadar böyle devam eder.Böylece bu safhada da sisteme ısı girişi devam etmiş olur.• Genleşme (4-5)• Artık silindire yakıt püskürtülmemektedir ve yanma durmuştur. Piston aşağıdoğru hareketine devam ettiğinden silindirdeki basınç da düşmeye başlar.• Egzoz (5-6)• Sistem 5 nolu noktaya (AÖN) geldiğinde egzoz valfi açılır. Silindir egzozsisitemi ile dışarıya açıldığından silindirdeki basınç atmosferik basınca düşer.Sistemden ısının atılması bu safhada gösterilmiştir. Gerçekte, dışarıya ısınınatılması pistonun egzoz stroğunu yapmasıyla olur (grafikte yatay çizgiylegösterilen strok), ancak ideal bir çevrimde egzoz stroğunda negatif veyapozitif bir iş yapılmadığından çevrimde incelenmez, ısının atılması da egzozvalfi açıldığında bir anda olmuş gibi gösterilir.

Karma Çevrim (Seilinger)

Summary


The efficiency of the ideal constant-volume combustion cycle is calculated in thesame manner as all thermal efficiencies:


Real internal-combustion engines do not operate according to ideal cycles, butrather with real gas, and are therefore subject to fluid, thermodynamic andmechanical losses.

Efficiency sequence (DIN 1940)The overall efficiency ηe includes the sum of all losses, and canthus be defined as the ratio of effective mechanical work to themechanical work equivalent of the supplied fuel:

WhereWe is the effective work available at the clutch andWB is the work equivalent of the supplied fuel.In order to better distinguish among the different losses, a further distinction canbe made:


The overall efficiency ηe includes the sum of all losses, and can thus be defined asthe ratio of effective mechanical work to the mechanical work equivalent of thesupplied fuel:

There are no operating conditions in which complete combustion takes place.A portion of the supplied fuel does not burn (hydrocarbon constituents in theexhaust gas), or fails to combust completely (CO in exhaust).ηB is often defined as "1" for small diesel engines at operating temperature and forcomparisons.


The efficiency index ηi is the ratio of indicated high-pressure work to the calorificcontent of the supplied fuel ηi = Wi/WB.The efficiency of cycle factor ηg includes all internal losses occurring in both high-pressure and low-pressure processes. These stem from:Real working gas, residual gas, wall heat losses, gas losses and pumping losses.For this reason, ηg is more appropriately broken down into ηgHD for the high-pressure portion and ηgLW for gas-exchange processes. The efficiency of cyclefactor therefore indicates how closely engine performance approaches thetheoretical ideal combustion cycle:ηg = ηgHD · ηgLW = Wi/Wth whereWi is the indicated work andWth is the work generated in the ideal-combustion cycle.Mechanical efficiency ηm defines the relationship between mechanical losses –especially friction losses in the crankshaft assembly and induction/exhaustsystems, and in oil and water pumps, fuel pump, alternator, etc. – and the workindex:ηm = We/Wi whereWe is the effective work available at the clutch andWi is the work index.The efficiency chain therefore appears as follows:ηe = ηB · ηth · ηgHD · ηgLW · ηm

YARARLANILAN KAYNAKLAR

Bosch Automotive Handbook, 2002

T.K. Garrett, K. Newton, “The Motor Vehicle”, Thirteenth Edition, 2001

http://hyperphysics.phy-astr.gsu.edu/hbase/thermo/diesel.html

Thermodynamic Ideal Cycle Analysis, http://www.springer.com/978-1-4614-3531-0

Yrd. Doç. Dr. Alp Tekin Ergenç, Motorlar Ders Notları, Yıldız Teknik

Yrd. Doç. Dr. Abdullah Demir, Sıkıştırma ile Ateşlemeli Motorlar sunum Notları, 2012

Documents

DieselDieselEngine Engine -CyclesCycles€¦ · Qianfan Xin, Diesel engine system design, Woodhead Publishing Limited, 2011 Diesel engineclassification. In external combustion, the