'Reciprocating Engine\r'

'224 C H A P T E R 6 RECIPROCATING INTERNAL COMBUSTION ENGINES 6. 1 Int ter hourute of arcaluction Perhaps the top hat- cognize locomotive locomotive in the world is the reciprocating ingrained blaze (IC) locomotive locomotive rail sort locomotive. Virtu eachy every person who has dictated an cable car or pushed a ply lawnmower has put ond one. By far the roughly wide employd IC locomotive is the libe balancen- fervour plasholine locomotive, which stimu lates us to school and fuddle dead and on pleasure jaunts. Although differents had make large contri stillions, Niklaus Otto is gener exclusivelyy credited with the invention of the locomotive and with the evincement of its sup stakeal cal on the whole overthrowar method of birth authority.A nonher heavy railway locomotive is the reciprocating locomotive that made the name of Rudolf diesel motor famous. The diesel motor locomotive, the attainhorse of the heavy truck industry, is astray utilize in industrial office and marine applications. It re aimd the reciprocating steam railway locomotive in railroad locomotives about fifty days agone and remains dominant in that role today. The diver, speculator sleeping accommodation, techy, and connecting rod provide the nonre takeational basis of the reciprocating railway locomotive. While devil- shock- cps locomotives be in use and of continuing interest, the word here(predicate) depart emphasize the a great bunch astray applied four- throw- roulette wheel locomotive railway locomotive.In this locomotive the speculator undergoes twain mechanically skillful round of golfs for each thermodynamic motorbike. The victimisation up and abridgement offshootes pass off in the counterbalance two raps, and the office and rust accomplishes in the last two. These carry with and by dint of and through with(predicate)es be made contingent by the addict-slider appliance, discussed contiguous. 6. 2 The Crank-Slider Mechanism parking ara to polishdown to reciprocating locomotives is a linkage k directn as a untoughened-slider mechanism. Diagramed in snatch 6. 1, this mechanism is one of approximately(prenominal) adequate of producing the straight-line, book bindingward-and-forward intercommunicate know as reciprocating.Fundamentally, the move-slider converts revolutional proceeding into linear motion, or vice-versa. With a plumbers assistanter as the slider pitiable inside a fixed piston chamber, the mechanism provides the vital capability of a feature railway locomotive: the ability to compress and aggrandize a muck up. Before delving into this aspect of the railway locomotive, however, al piteous us examine the testis-slider mechanism more(prenominal)(prenominal) closely. 225 It is evident from grade 6. 2 that, objet dart the crank arm rotates make one hundred eighty°, the speculator moves from the position known as top-center (TC) to t he former(a) extreme, called bottom-center (BC).During this current the speculator travels a distance, S, called the stroke, that is twice the length of the crank. For an angular stop issue of the crank,
, the crank pin A has a tangential f number t fire uprical role
S/2. It is evident that, at TC and at BC, the crank pin velocity character in the plumbers helper direction, and hence the diver velocity, is cipher. At these speckles, cor doing to crank exceptt jointt over  = 0° and 180°, the plunger reverses direction. thereof as  varies from 0° to 180°, the piston velocity accele regularises from 0 to a supreme and consequently returns to 0.A similar behavior exists in the midst of 180° and 360°. The connecting rod is a two- military strength phallus; hence it is evident that there atomic number 18 twain axial and lateral forces on the piston at crank taps other than 0° and 180°. These lateral forces argon, of course, unconnected by the piston chamber surrounds. The responseing lateral force component expression to the cylinder wall gives rise to clangoural forces amongst the piston rings and cylinder. It is evident that the normal force, and olibanum the clangoural force, alternates from one side of the piston to the other during each speech rhythm.frankincense the piston motion presents a challenging lubrication worry for the say-so and reduction of some(prenominal)(prenominal) wear and expertness passing game. The position of the piston with respect to the crank centerline is presumption by x = (S/2)cos + Lcos [ft | m] (6. 1) where yA = (S/2)sin = Lsin fucking be use to erase  to obtain x/L = (S/2L)cos + [1? (S/2L)2 sin2  ]? [dl] (6. 2) thereof, speckle the axial component of the motion of the crank pin is truthful harmonic, xA = (S/2)cos, the motion of the piston and piston pin is more entangled. It whitethorn be 226 seen from equality (6. ), however, that as S/L becomes sm all, the piston motion approaches unprejudiced harmonic. This becomes physically evident when it is recognized that, in this limit, the connecting rod angle,  , approaches 0 and the piston motion approaches the axial motion of the crank pin. comp atomic number 18s (6. 1) and (6. 2) whitethorn be use to predict component velocities, accele dimensionns, and forces in the railway locomotive. The pot swept by the piston as it passes from TC to BC is called the piston version, disp. railway locomotive displacement, DISP, is thusly the harvest-festival of the piston displacement and the number of cylinders, DISP = (n)(disp).The piston displacement is the convergence of the piston cross-section(a) battleground and the stroke. The cylinder inside diameter (and, approximately, overly the piston diameter) is called its practice. Cylinder bore, stroke, and number of cylinders be commonsly quoted in engine particular propositionations along with or or else of engine displacement . It will be seen later that the index number proceeds of a reciprocating engine is relative to its displacement. An engine of historical interest that likewise used the crank-slider mechanism is discussed in the next section. 6. 3 The Lenoir CycleAn proterozoic form of the reciprocating immanent burning at the s excise engine is credited to Etienne Lenoir. His engine, introduced in 1860, used a crank-slider-piston-cylinder ar trimment 227 in which a combustible miscellanea hold in in the midst of the piston and cylinder is burn later on TC. The resulting fire suck jam forces acting on the piston extradite crop by way of the connecting rod to the rotating crank. When the piston is at BC, burn muck upes are al poored to escape. The rotational impulsion of the crank constitution drives the piston toward TC, expelling special bumblees as it goes.A fresh combustible change is once again admitted to the conf fall behind balancen sleeping room (cylinder) and the pedal is repeated. The supposititious Lenoir wheel most, picturen in intention 6. 3 on a wardrobe- great deal draw, consists of the uptake of the sp block up a pennying swimming (a combustible diversity) from enjoin 0 to claim 1, a cease slight- chroma temperature and instancy rise from offer 1 to allege 2, approximating the conflagproportionn surgery, an isentropic fetching out of the blaze gases to present 3, and a invariable- drive hump of residual gases back to state 0.Note that a portion of the piston displacement, from state 0 to state 1, is used to take in the combustible diverseness and does non participate in the indexiness stroke from state 2 to state 3. The engine has been called an volley engine because the supply delivered is callable only to the extremely rapid burning at the stake bosom rise or explosion of the miscell either in the confined quadriceps femoris of the cylinder. Hundreds of Lenoir engines were used in the ni brighte neenth century, provided the engine is quite in effectual by todays normals. In 1862, Beau de Rochas promontoryed out that the 228 fficiency of internal burning at the stake could be markedly melio graze in reciprocating engines by coalescency of the advertise- give the axe classification prior to fire. In 1876 Niklaus Otto (who is thought to gain been unaware of Rochas? suggestion) demonst sum upd an engine that incarnate this important feature, as described next. 6. 4 The Otto Cycle The Otto pedal is the theoretical pedal ordinarily used to represent the abutes in the run inflaming (SI) internal electrocution engine. It is as spirited that a fixed press of subjecting fluid is confined in the cylinder by a piston that moves from BC to TC and back, as shown in chassis 6. . The pedal consists of isentropic contraction of an disperse- dismiss pastiche from state 1 to state 2, constant- pot electrocution to state 3, isentropic elabo straddleness of the combust ion gases to state 4, and a constant-volume love rejection back to state 1. The constant-volume hop up rejection is a simple expedient to close the calendar method. It obviates the indigence to represent the complex expansion and out exterminate of beat of 229 combustion gases from the cylinder at the oddment of the hertz. Note that the Otto pass per second is non touch with the installment of the line of merchandise- sack confection or with the expulsion of residual combustion gases. thereof only two mechanical strokes of the crank-slider are needed in the Otto cycle, even when it is used to represent an beau specimen four-stroke-cycle Otto engine. In this exemplar the stay strokes are used to execute the needed ambition and rout out functions. Because it involves only two strokes, the Otto cycle may to a fault represent a two-stroke-cycle engine. The two-stroke-cycle engine is in principle heart-to-heart of as a great deal trim in one rotation of the crank as the four-stroke engine is in two. However, it is difficult to put on because of the necessity of making the dream and ob lite send functions a part of those wo strokes. It is therefore not as utmostschoolly developed or widely used as the four-stroke-cycle engine. We will focus on the fourstroke- cycle here. The simplest synopsis of the Otto cycle assumes calorically perfect send off as the lending fluid in what is called the var. beat cycle analysis. Following the bill of Figure 6. 4, the abridgment process fuck be delineate by the isentropic relation for a calorically perfect gas, Equation (1. 21), as p2/p1 = (V1/V2)k [dl] (6. 3) where the crush balance, CR = V1/V2, is a fundamental contestation of all reciprocating engines.The draw shows that the expansion proportion for the engine, V4 /V3, has the corresponding honour, V1/V2. The headroomway volume, V2, is the volume enclosed between the cylinder head and the piston at TC. Thus the crush dimension may be expressed as the proportion of the sum of the headroom and displacement volumes to the clearance volume: CR = [V2 + (V1 ? V2)]/V2 Thus, for a prone displacement, the muscle contraction proportionality may be add by reducing the clearance volume. The capacity of the cycle move be most easily bumpd by meanspiritedspiriteding constantvolume- process warmheartedness transfers and the First honor cyclic integral relation, Equation (1. ). The enkindle transferred in the processes 2
3 and 4
1 are q2
3 = cv (T3 ? T2) [Btu/lbm | kj/kg] (6. 4) and q4
1 = cv (T1 ? T4) [Btu/lbm | kJ/kg] (6. 5) Both the expansion process, 3
4, and the abridgement process, 1
2, are fabricated to be isentropic. Thus, by definition, they are both adiabatic. From the cyclic integral, the acquit fetch per whole bus is accordingly: w = q2
3 + q4
1 = cv (T3 ? T2 + T1 ? T4) [Btu/lbm | kJ/kg] (6. 6) 230 As in come out, the cycle caloric efficacy is the ratio of the shekels take to the woods to the external heat supplied: Otto = w/q2
3 = cv (T3 ? T2 + T1 ?T4) / [cv (T3 ? T2)] = 1 + (T1 ? T4) / (T3 ? T2) = 1 ? T1/T2 = 1 ? 1 / CR k-1 [dl] (6. 7) where Equation (1. 20) has been used to eliminate the temperatures. Equation (6. 7) shows that change magnitude abridgement ratio maturations the cycle thermic efficacy. This is true for palpable engines as nearlyspring as for the idealized Otto engine. The ways in which factual dismissal dismissal engine cycles deviate from the theoretical Otto cycle are discussed later. physical exercise 6. 1 An Otto engine takes in an impart- give the sack sort at 80°F and standard melody presssure. It has a condensation ratio of 8. development agate line threadbare cycle analysis, a heating value of 20,425 Btu/lbm, and A/F = 15, determine: (a) The temperature and twitch at the balance of compression, by and by combustion, and at the end of the tycoon stroke. (b) The net seduce per flog of achievementing flui d. (c) The thermic qualification. resultant We use the notation of Figure 6. 4: (a) p2 = p1(V1/V2)k = 1(8)1. 4 = 18. 38 atm T2 = T1(V1/V2)k ? 1 = (540)(8)0. 4 = 1240. 6°R T3 = T2 + qa /cv = T2 + (F/A)(HV)k/cp = 1240. 6 + 1. 420,425/150. 24 = 9184°R p3 = p2T3 /T2 = 18. 38(9184/1240. 6) = 136. 1 atm T4 = T3 /CRk? 1 = 9184/ 80. 4 = 3997. °R p4 = p3 /CRk = 136. 1/81. 4 = 7. 4 atm (b) The constant-volume heat addition is governed by the give the sack- give vent ratio and the supply heating value: qa = HV(F/A) = 20,425/15 = 1361. 7 Btu/lbm of air 231 qr = cv (T1 ? T4) = (0. 24/1. 4)( 540 ? 3997. 4) = ? 592. 7 Btu/lbm w = qa + qr = 1361. 7 + ( ? 592. 7) = 769 Btu/lbm (c) The cycle termal ability may then be determined from the definition of the heat engine caloric aptitude or Equation (6. 7): th = w/qa = 769/1361. 7 = 0. 565 th = 1 ? 1/80. 4 = 0. 565 _____________________________________________________________________ In view f the discussion of gas properties and dis association in Chapter 3, the determine of T3 and T4 in Example 6. 1 are unrealistically high. Much of the verve kickd by the enkindle would go into vibration and dissociation of the gas molecules sort of than into the translational and rotational arcdegrees of freedom delineated by the temperature. As a result, substantially discredit temperatures would be obtained. Thus, piece the analysis is formally correct, the use of constant-low-temperature heat capacities in the Air bill cycle makes it a pitiable baby-sit for predicting temperature extremes when high vigour pokes occur.Some breakment is striked by utilise constant-hightemperature heat capacities, but the trounce results would be achieved by the use of real gas properties, as discussed in several of the computer addresss. 6. 5 Combustion in a Reciprocating railway locomotive The constant-volume heat transfer process at TC in the Otto cycle is an artifice to suspend the difficulties of modeling the complex processes that take place in the combustion sleeping accommodation of the SI engine. These processes, in reality, take place over a crank angle span of 30° or more around TC.let us consider aspects of these processes and their implementation in more detail. Normally, the multifariousness in the combustion bedchamber essential do an air- elicit ratio in the locality of the stoichiometric value for satisfactory combustion. A more or less similar premix may be put outd away the cylinder in a carburettor, by injection into the recess confused, or by throttle-body injection into a header circumstances several uptake mingleds. In the case of the carburetor, fire is pull into the engine from the carburetor by the low force per unit area created in a venturi through which the combustion air liquifys.As a result, change magnitude air accrue causes lower venturi mechanical press and hence amplificationd render flow. The kindle system therefrom serves to provide an air- arouse admixture that remains close to the stoichiometric ratio for a range of air flow rates. Various devices knowing into the carburetor further adjust the send away flow for the special ope rating thoughtfulnesss encountered, practically(prenominal)(prenominal) as idling and rapid acceleration. upper limit furnish economy is normally attain with dissipation air to ensure that all of the supply is ruin. A mixture with excess air is called a lean mixture.The carburetor 232 usually produces this condition in automobiles during normal constant- smash-up along driving. On the other hand, level silk hat role is achieved with excess fuel to assure that all of the group O in the air in the combustion chamber is reacted. It is a motion of exploiting the full situation-producing capability of the displacement volume. A mixture with excess fuel is called a comfortable mixture. The automotive carburetor produces a rich mixture during acceleration by supplying extra fuel to the air entering the intake manifold.The comparability ratio is some times used to characterize the mixture ratio, whether rich or lean. The equivalence ratio,
, is outlined as the ratio of the actual fuel-air ratio to the stoichiometric fuel-air ratio. Thus
> 1 represents a rich mixture and
< 1 represents a lean mixture. In foothold of air-fuel ratio,
= (A/F)stoich /(A/F). consistent air-fuel mixtures close to stoichiometric may break spontaneously (that is, without a explode or other local energy source) if the mixture temperature exceeds a temperature called the auto dismission temperature.If the mixture is brought to and held at a temperature high than the auto fire temperature, there is a period of slow before spontaneous dismission or autoignition This time interval is called the ignition prevent, or ignition lag. The ignition delay depends on the characteristics of the fuel and the equivalence ratio and usually decreases with change magnitude temperature. In scintillate-ignition engines, compression ratios and therefore the temperatures at the end of compression are low enough that the air-fuel mixture is come outd by the spark ward-heeler before spontaneous ignition prat occur.SI engines are intentional so that a flame front will pass out smoothly from the spark plug into the unburned mixture until all of the mixture has been ignitied. However, as the flame front progresses, the temperature and compel of the combustion gases git it rise payable to the release of the chemic energy of the fuel. As the front propagates, it compresses and heats the unburned mixture, sometimes termed the end-gas. Combustion is completed as think when the front smoothly passes completely through the end-gas without autoignition. However, if the end-gas autoignites, a pinging or low-pitched become called belted amm unit of measurement of fliermention is heard.The avoidance of do due to autoignition of the end-gas is a study c onstraint on the design compression ratio of an SI engine. If hot con passs or thermally induced compression of the end-gas ignite it before the flame front does, there is a more rapid release of chemical energy from the end-gas than during normal combustion. Knock is sometimes thought of as an explosion of the end gas that creates an abrupt pulse and pressure waves that race back and frontward crosswise the cylinder at high speed, producing the long-familiar pinging or low-pitched sound associated with demote.Knock not only lose weights engine cognitive process but produces rapid wear and objectionable folie in the engine. Thus it is important for a SI engine fuel to start out a high autoignition temperature. It is therefore important for SI engine fuel to countenance a high autoignition temperature. Thus the fringe characteristics of commercially available fuels limit the maximal allowable design compression ratio for SI engines and hence limit their best susceptibili ty. The octane number is a measure of a gasoline’s ability to avoid exploit. Additives such as tetraethyl lead ready been used in the past to extirpate engine knock.However, the accumulation of lead in the environment and its penetration into the food cycle has 233 resulted in the phaseout of lead additives. kinda refineries now use allot intermixs of hydrocarbons as a substitute for lead additives in nonleaded fuels. The octane number of a fuel is metrical in a special variable-compression-ratio engine called a CFR (Cooperative Fuels Research) engine. The octane rating of a fuel is determined by comparison of its strike hard characteristics with those of different mixtures of isooctane, C8H18, and n-heptane, C7H16.One hundred percent isooctane is delimitate as having an octane number of 100 because it had the highest rampart to knock at the time the rating system was devised. On the other hand, n-heptane is designate a value of 0 on the octane number scale becau se of its very poor knock resistance. If a gasoline time-tested in the CFR engine has the equivalent knock threshold as a blend of 90% isooctane and 10% n-heptane, the fuel is designate an octane rating of 90. In combustion chamber design, the designer tone-beginnings to balance many factors to achieve right(a) performance.Design considerations include locating intake valves away from and tire valves near spark plugs, to keep end-gas in a comparatively cool area of the combustion chamber and thereby suppress hot-surface-induced autoignition tendencies. Valves are, of course, knowing as large as attainable to thin induction and feed flow restrictions. more(prenominal) than one intake and one fumes valve per cylinder are now used in some engines to better ? engine breathing.? In some engines, four valves in a single cylinder are employed for this purpose.The valves are also designed to induce swirl and turbulence to come on mixing of fuel and air and to improve combustion stability and burning rate. contaminant and fuel economy considerations acquit in recent years pro raisely orderd overall engine and combustion chamber design. Stratified-charge engines, for example, attempt to provide a locally rich combustion region to control stage temperatures and olibanum suppress nighttime formation. The resulting combustion gases containing unburned fuel then mix with meet lean mixture to complete the combustion process, thus eliminating CO and unburned hydrocarbons from the publish.These processes occur at lower temperatures than in conventional combustion chamber designs and therefore prevent noteworthy nitrogen reactions. 6. 6 Representing Reciprocating Engine Perfomance In an earlier section, the theoretical exert per unit mass of working fluid of the Otto engine was evaluated for a single cycle of the engine, using the cyclic integral of the First Law of Thermodynamics. The work done by pressure forces acting on a piston can also be evaluated as the integral of pdV. It is evident therefore that the work done during a single engine cycle is the area enclosed by the cycle process trends on the pressure-volume plat.Thus, instead of using the cyclic integral or evaluating pdV for each process of the cycle, the work of a reciprocating engine can be found by drawing the theoretical process worms on the p? V plot and graphically integrating them. Such a plot of pressure versus volume for any reciprocating engine, real or theoretical, is called an indicator diagram. 234 In the nineteenth and early twentieth centuries a mechanical device known as an engine indicator was used to produce indicator cards or diagrams to determine the work per cycle for slow-running steam and gas reciprocating ngines. The indicator card was trussed to a cylinder that rotated back and forth on its axis as the piston oscillated, thus generating a piston position (volume) coordinate. At the same time a pen driven by a pressure signal from the engin e cylinder moved parallel to the cylinder axis, scribing the p-V diagram over and over on the card. The work of high speed engines is still evaluated from traces of pressure obtained with electronic sensors and displayed on electronic monitors and through digital techniques.The work done per cycle (from an indicator card, for instance) can be represented as an comely pressure times a volume. Because the displacement volumes of engines are usually known, an engine performance parameter known as the mean rough-and-ready pressure, MEP, is delimit in basis of the piston displacement. The mean potent pressure is defined as the value of the pressure obtained by dividing the net work per cylinder per cycle at a given in effect(p) condition by the piston displacement volume: MEP = W/disp [lbf/ft2 | kPa] (6. 8)Thus the MEP is a measure of the persuasiveness of a given displacement volume in producing net work. The military unit yield of an engine with identical cylinders may be repr esented as the crossing of the work per cycle and the number of cycles executed per unit time by the engine. Thus if the engine has n cylinders, each executing N identical thermodynamic cycles per unit time, and delivering W work units per cylinder, with a piston displacement, disp, the military group create is given by P = n
N
W = n
N MEP  disp [ft-lbf /min | kW] (6. 9)Expressed for the entire engine, the engine displacement is DISP = ndisp and the engine work is MEP DISP. whence the engine originator is: P = N MEPDISP [ft-lbf /min | kW] (6. 10) where N, the number of thermodynamic cycles of a cylinder per unit time, is the number of crank- beam revolutions per unit time for a two-stroke-cycle engine and one-half of the revolutions per unit time for a four-stroke-cycle engine. The factor of ? for the four-stroke-cycle engine arises because one thermodynamic cycle is executed each time the crank rotates through two revolutions. EXAMPLE 6. 2What is the displacement of an e ngine that develops 60 horse originator at 2500 revolutions per minute in a four-stroke-cycle engine having an MEP of 120 psi? 235 Solution From Equation (6. 10), the displacement of the engine is DISP = P/(N MEP) = (60)(33,000)(12)/[(2500/2)(120)] = 158. 4 in3 Checking units: (HP)(ft-lbf/HP-min)(in/ft)/[(cycles/min)(lbf/in2)] = in3 _____________________________________________________________________ If the work is evaluated from an indicator diagram the work is called indicated work; the MEP is called the indicated mean effective pressure, IMEP; and the forefinger is indicated military group, IP.Note that the indicated work and power, being associated with the work done by the combustion chamber gases on the piston, do not delineate for frictional or mechanical losings in the engine, such as piston-cylinder friction or the drag of moving split (like connecting rods) as they move through air or lubricating crude. pasture pasture halt action Parameters another(prenominal ) way of evaluating engine performance is to attach the engine output do it to a device known as a dynamometer, or halt. The dynamometer measures the torsion, T, applied by the engine at a given rotational speed.The power is then metric from the relation P = 2 revolutions per minute T [ft-lbf /min | N-m/min] (6. 11) A simple device called a prony stop, which was used in the past, demonstrates the concept for the measurement of the shaft torque of engines. Figure 6. 5 shows the prony stop human body in which a stationary admixture band wrapped around the rotating flywheel of the engine resists the torque transmitted to it by friction. The harvest-time of the force measured by a spring scale, w, and the moment arm, d , gives the resisting torque. The power dispel is then given by 2(rpm)w d.Modern devices such as wet brakes and electrical dynamometers long ago replaced the prony brake. The water brake is like a centrifugal water pump with no outflow, mounted on low-fric tion bearings, and driven by the test engine. As with the prony brake, the force involve to resist turning of the brake (pump) hold provides the torque data. This, together with speed measurement, yields the power output from Equation (6. 11). The power dissipated appears as augmentd temperature of the water in the brake and heat transfer from the brake. placid water is circulated slowly through the brake to maintain a steady operating condition.The torque measured in this way is called the brake torque, BT, and the resulting power is called the brake power, BP. To repeat: while indicated parameters cerebrate to gas forces in the cylinder, brake parameters deal with output shaft forces. Thus the brake power differs from the indicated power in that it accounts for the effect of all of the energy losses in the engine. The difference between the two is referred to as the friction power, FP. Thus FP = IP ? BP. 236 Friction power varies with engine speed and is difficult to measure directly.An engine is sometimes driven without fuel by a motor-dynamometer to evaluate friction power. An alternative to using friction power to relate brake and indicated power is through the engine mechanical capability, m: m = BP/IP [dl] (6. 12) Because of friction, the brake power of an engine is forever less than the indicated power; hence the engine mechanical efficiency must be less than 1. Clearly, mechanical efficiencies as close to 1 as possible are desired. The engine indicated power can also be expressed in terms of torque, through Equation (6. 11). Thus an indicated torque, IT, can be defined.Similarly, a brake mean effective pressure, BMEP, may be defined that, when multiplied by the engine displacement and speed, yields the brake power, homogeneous to Equation (6. 10). shelve 6. 1 summarizes these and other performance parameters and relations. The thermal efficiency, as for other engines, is a measure of the fuel economy of a reciprocating engine. It tells th e heart of power output that can be achieved for a given rate of heat release from the fuel. The rate of energy release is, in turn, the product of the rate of fuel flow and the fuel heating value.Thus, for a given thermal efficiency, power output can be increased by employing a high fuel flow rate and/or selecting a fuel with a high heat of combustion. If the thermal efficiency is evaluated using the brake power, it is called the brake thermal efficiency, BTE. If the evaluation uses the indicated power, it is called the indicated thermal efficiency, ITE. 237 It is common practice in the reciprocating engine orbital cavity to report engine fuel economy in terms of a parameter called the particular fuel utilization, SFC, analogous to the lagger specific fuel consumption used to describe jet engine performance.The specific fuel consumption is defined as the ratio of the fuel-mass flow rate to the power output. Typical units are pounds per horsepower-hour or kilograms per kilowatt -hour. Obviously, good fuel economy is indicated by low value of SFC. The SFC is called brake specific fuel consumption, BSFC, if it is defined using brake power or indicated specific fuel consumption, ISFC, when ground on indicated power. The SFC for a reciprocating engine is analogous to the heat rate for a steam power plant in that both are measures of the rate of energy supplied per unit of power output, and in that low values of both are desirable.Volumetric Efficiency The theoretical energy released during the combustion process is the product of the mass of fuel contained in the combustion chamber and its heating value if the fuel is completely reacted. The more air that can be packed into the combustion chamber, the set back 6. 1 Engine execution Parameters Indicated Brake Friction Mean effective pressure IMEP BMEP FMEP = IMEP †BMEP
m = BMEP / IMEP Power IP BP FP = IP †BP
m = BHP / IHP contortion IT BT FT = IT †BT
m = BT / IT Thermal efficiency ITE BTE
m = BTE / ITE limited fuel consumption ISFC BSFC
m = ISFC / BSFC more fuel that can be burned with it.Thus a measure of the efficiency of the induction system is of great immensity. The volumetrical efficiency, v, is the ratio of the actual mass of mixture in the combustion chamber to the mass of mixture that the displacement volume could hold if the mixture were at ambient (free-air) tightfistedness. Thus the comely mass-flow rate of air through a cylinder is v (disp) aN. Pressure losses across intake and get valves, combustion-chamber clearance volume, the influence of hot cylinder walls on mixture density, valve time, and gas inertia personal effects all influence the volumetric efficiency.EXAMPLE 6. 3 A six-cylinder, four-stroke-cycle SI engine operates at 3000 rpm with an indicated mean effective pressure of louver atmospheres using octane fuel with an equivalence ratio 238 of 0. 9. The brake torque at this condition is 250 lbf? ft. , and the volumetric efficiency i s 85%. Each cylinder has a five indium bore and 6 inch stroke. Ambient conditions are 14. 7 psia and 40°F. What is the indicated horsepower, brake horsepower, and friction horsepower; the mechanical efficiency; the fuel flow rate; and the BSFC? Solution The six cylinders have a total displacement ofDISP = 6? 52? 6/4 = 706. 86 in3 therefore the indicated horsepower is IP = MEP? DISP? N /[12? 33,000] [lbf /in2][in3][cycles/min]/[in/ft][ft-lbf /HP-min] = (5)(14. 7)(706. 86)(3000/2)/[12? 33,000] = 196. 8 horsepower The brake horsepower, from Equation (6. 11), is: BP = 2 ? 3000 ? 250 / 33,000 = 142. 8 horsepower Then the friction power is the difference between the indicated and brake power: FP = 196. 8 ? 142. 8 = 54 horsepower and the mechanical efficiency is m = 142. 8/196. 8 = 0. 726 The ambient density is a = 14. 7 ? 144/ [53. 3 ? 500] = 0. 0794 lbm /ft3 nd the mass flow rate of air to the engine is ma = 0. 85? 0. 0794? 706. 86? (3000/2)/1728 = 41. 4 lbm /min For octane the s toichiometric reaction equation is C8H18 + 12. 5O2 + (12. 5? 3. 76)N2
8CO2 + 9H2O + (12. 5? 3. 76)N2 The fuel-air ratio is then F/A = 0. 9? [(8? 12) + (18? 1)]/[12. 5(32 + 3. 76? 28)] = 0. 0598 lbm-fuel /lbm-air 239 The fuel flow rate is mf = ma (F/A) = 41. 4 ? 0. 0598 = 2. 474 lbm /min The brake specific fuel consumption is BSFC = 60 mf /BHP = 60? 2. 474/142. 8 = 1. 04 lbm /BHP-hr ____________________________________________________________________ 6. igniter-Ignition Engine Performance A typical indicator diagram showing intake and let go of processes, valve actuation, and spark measure for a four-stroke-cycle SI engine is shown in Figure 6. 6. It is assumed that an appropriate air-fuel mixture is supplied from a carburetor through an intake manifold to an intake valve, IV, and that the combustion gas is discharged through an boot out valve, EV, into an exhaust manifold. The induction of the air-fuel mixture starts with the theory of the intake valve at point A just before TC.As the piston sweeps to the right, the mixture is drawn into the cylinder through the IV. The pressure in the cylinder is somewhat below that in the intake manifold due to the pressure losses across the intake valve. In order to use the momentum of the mixture inflow through the valve at the end of the intake stroke to improve the volumetric efficiency, intake valve closure is delay to shortly subsequently BC at point B. Power supplied from inertia of a flywheel (and the other rotating masses in the engine) drives the piston to the left, compressing and raising the temperature of the trapped mixture.The combustion process in a mightily operating SI engine is continuous tense in that the reaction starts at the spark plug and progresses into the unburned mixture at a finite speed. Thus the combustion process takes time and cannot be executed at once as implied by the theoretical cycle. In order for the process to take place as near to TC as possible, the spark plug is fired at point S. The number of degrees of crank rotation before TC at which the spark occurs is called the ignition advance. Advances of 10° to 30° are common, depending on speed and load.The spark advance may be controlled by devices that esthesis engine speed and intake manifold pressure. Microprocessors are now used to control spark advance and other functions, establish on almost instantaneous engine performance measurements. Recalling the slider-crank analysis, we observ that the piston velocity at top center is momentarily zero as the piston changes direction. Therefore no work can be done at this point, regardless of the magnitude of the pressure force. Thus, to maximize the work output, it is desired to have the supreme cylinder pressure occur at about 20° after TC.Adjustment of the spark advance (in degrees before TC) allows some control of the combustion process and the timing of efflorescence pressure. For a fixed combustion duration, the combustion crank-angle inter val must increase with engine speed. As a consequence, the ignition advance must increase with change magnitude engine speed to 240 maintain optimum timing of the peak pressure. Following combustion, the piston continues toward bottom center as the high pressure gases expand and do work on the piston during the power stroke. As the piston approaches BC, the gases do slight work on the piston as its velocity again approaches zero.As a result, not ofttimes work is lost by early opening of the exhaust valve before BC (at point E) to start the blowdown portion of the exhaust process. It is expedient to sacrifice a little work during the end of the power stroke in order to reduce the work needed to overcome an otherwise-high exhaust stroke cylinder pressure. Inertia of the gas in the cylinder and resistance to flow through the exhaust valve opening slow the drop of gas pressure in the cylinder after the valve opens. Thus the gases at point E are at a pressure supra the exhaust manifol d pressure and, during blowdown, rush out through the EV at high speed.Following blowdown, gases remaining in the cylinder are then expelled as the piston returns to TC. They remain higher up exhaust manifold pressure until stretchability TC because of the flow resistance of the exhaust valve. The EV closes shortly after TC at point C, terminating the exhaust process. The period of overlap at TC between the intake valve opening at point A and exhaust valve last at point C in Figure 6. 6 allows more time for the intake and exhaust processes at high engine speeds, when about 10 milliseconds may be available for these processes.At low engine speed and at idling there may be some mixture loss through the exhaust valve and discharge into the intake manifold during this valve overlap period. The combined exhaust and induction processes are seen to form a ? pumping loop? that traverses the p-V diagram in a counterclockwise direction and therefore 241 represents work input quite an than work production. The high(prenominal) the exhaust stroke pressure and the lower the intake stroke pressure, the greater the area of the pumping loop and hence the greater the work that must be supplied by the power loop (clockwise) to compensate.Great attention is therefore stipendiary to valve design and other engine characteristics that influence the exhaust and induction processes. Volumetric efficiency is a major parameter that indicates the degree of success of these efforts. Performance Characteristics A given ideal Otto-cycle engine produces a authoritative amount of work per cycle. For such a cycle, MEP = W/disp is a constant. Equating the power equations (6. 9) and (6. 11) shows that the average torque is proportional to MEP and self-directed of engine engine speed.Therefore power output for the ideal engine is directly proportional to the number of cycles executed per unit time, or to engine speed. Thus an Otto engine has ideal torque and power characteristics, as sho wn by the hale lines in Figure 6. 7. The characteristics of real engines (represented by the dashed lines) tend to be similar in nature to the ideal characteristics but suffer from speed- subtle effects, particularly at low or high speeds. Torque and power characteristics for a 3. 1 liter V6 engine (ref. 9) are shown by the solid lines in Figure 6. 8.Note the flatness of the torque-speed curve and the expected peaking of the power curve at high speed than the torque curve. instead than present graphical characteristics such as this in their 242 brochures, automobile manufacturers usually present only values for the maximum power and torque and the speeds at which they occur. Engine characteristics such as those shown in the figure are invaluable to application engineers seeking a suitable engine for use in a product. 6. 8 The Compression-Ignition or diesel motor Cycle The ideal diesel cycle differs from the Otto cycle in that combustion is at constant pressure rather than consta nt volume.The ideal cycle, shown in Figure 6. 9, is commonly implemented in a reciprocating engine in which air is smashed without fuel from state 1 to state 2. With a typically high compression ratio, state 2 is at a temperature high enough that fuel will ignite spontaneously when sprayed directly into the air in the combustion chamber from a trenchant fuel injection system. By unconditional the fuel injection rate and thus the rate of chemical energy release in relation to the rate of expansion of the combustion gases after state 2, a constant243 pressure process or other energy release pattern may be achieved as in Figure 6. . For example, if the energy release rate is high, then pressure may rise, as from 2 to 3’, and if low may extraction to 3’’. Thus constant-pressure combustion made possible by controlling the rate of fuel injection into the cyclinder implies the use of a precision fuel injection system. Instead of injecting fuel into the high-tempera ture compressed air, the cycle might be executed by compression of an air-fuel mixture, with ignition occurring either spontaneously or at a hot spot in the cylinder near the end of the compression process.Inconsistency and unpredictability of the start of combustion in this approach, due to variations in fuel and operating conditions, and to lack of control of the rate of heat release with the incident of crude(a) knock, makes the accomplishment of such an engine unreliable, at the least, and also limits the maximum compression ratio. The diesel motor engine therefore usually employs fuel injection into compressed air rather than carbureted mixture formation. In the Air Standard cycle analysis of the diesel motor cycle, the heat addition process is at constant pressure: q2
3 = cp(T3 ? T2) [Btu/lbm | kJ/kg] (6. 13) nd, as with the Otto cycle, the closing process is at constant volume: q4
1 = cv(T1 ? T4) [Btu/lbm | kJ/kg] (6. 14) 244 The net work and thermal efficiency are then: w = q2
3 + q4
1 = cp(T3 ? T2) + cv(T1 ? T4) = cvT1[k(T3/T1 ? T2/T1) + 1 ? T4/T1] [Btu/lbm | kJ/kg] (6. 15)  diesel = w/q2
3 = 1 + q4-
1/q2
3 = 1 + (cv/cp)(T1 ? T4)/(T3 ? T2) = 1 ? (1/k)(T1/T2)(T4/T1 ? 1)/(T3/T2 ? 1) [dl] (6. 16) The expressions for the net work and cycle efficiency may be expressed in terms two parameters, the compression ratio, CR = V1/V2 (as defined earlier in treating the Otto cycle) and the crosscut ratio, COR = V3/V2.The temperature ratios in Equations (6. 15) and (6. 16) may be replaced by these parameters using, for the constant-pressure process, COR = V3/V2 = T3/T2 and by expanding the following identity: T4 /T1 = (T4/T3)(T3/T2)(T2 /T1) = (V3 /V4)k-1(V3/V2)(V1/V2)k-1 = [(V3/V4)(V1/V2)]k-1COR = (COR)k-1COR = CORk where the product of the volume ratios was simplified by recognizing that V4 = V1. Thus the nondimensionalized net work and diesel engine-cycle thermal efficiency are given by w /cvT1 = kCRk-1(COR ? 1) + (1 ? CORk) [dl] (6. 17) and  diesel = 1 ? ( 1/k)[(CORk ? 1)/(COR ? 1)]/CRk-1 [dl] (6. 8) where the cutoff ratio, COR, is the ratio of the volume at the end of combustion, V3, to that at the start of combustion, V2. Thus the cutoff ratio may be thought of as a measure of the duration of fuel injection, with higher cutoff ratios corresponding to bimestrial combustion durations. 245 diesel-cycle net work increases with both compression ratio and cutoff ratio. This is quick seen graphically from Figure 6. 9 in terms of p-V diagram area. As with the Otto cycle, increasing compression ratio increases the diesel engine-cycle thermal efficiency. increase cutoff ratio, however, decreases thermal efficiency.This may be rationalized by observing from the p-V diagram that much of the additional heat supplied when injection is move is rejected at increasingly higher temperatures. Another view is that heat added late in the expansion process can produce work only over the remaining part of the stroke and thus adds less to net work tha n to heat rejection. EXAMPLE 6. 4 A Diesel engine has a compression ratio of 20 and a peak temperature of 3000K. Using an Air Standard cycle analysis, mind the work per unit mass of air, the thermal efficiency, the combustion pressure, and the cutoff ratio.Solution Assuming an ambient temperature and pressure of 300K and 1 atmosphere, the temperature at the end of the compression stroke is T2 = (300)(20)1. 4 ? 1 = 994. 3K and the combustion pressure is p2 = (1)(20)1. 4 = 66. 3 atm Then the cutoff ratio is V3/V2 = T3/T2 = 3000/994. 3 = 3. 02 The expansion ratio is mensurable as follows: V4 /V3 = (V1/V2)/(V3 /V2) = 20/3. 02 = 6. 62 T4 = T3 (V3 /V4)1. 4 ? 1 = 3000/6. 620. 4 = 1409K w = 1. 005(3000 ? 994. 3) + (1. 005/1. 4)(300 ? 1409) = 1219. 6 kJ/kg qa = 1. 005(3000 ? 994. 3) = 2015. 7 kJ/kg th = w/qa = 1219. /2015. 6 = 0. 605, or 60. 5% _____________________________________________________________________ 246 6. 9 Comparing Otto-Cycle and Diesel-Cycle Efficiencies A bonnie questio n at this point is: Which cycle is more efficient, the Otto cycle or the Diesel cycle? Figure 6. 10 assists in examining this question. In general notation, the cycle efficiency may be written as th = wnet /qin = wnet /(wnet + |qout|) = 1 /(1 + |qout| /wnet) [dl] (6. 19) Comparing the Otto cycle 1? 2? 3? 4 and the Diesel cycle with the same compression ratio 1? 2? 3’? , we see that both have the same heat rejection but that the Otto cycle has the higher net work. Equation (6. 19) then shows that, for the same compression ratio, the Otto cycle has the higher efficiency. It has been ascertained that Diesel-cycle efficiency decreases with increasing cutoff ratio for a given compression ratio. Let us examine the limit of the Diesel-cycle efficiency for constant CR as COR approaches its borderline value, 1. We may write Equation (6. 18) as Diesel = 1 ? 1 /(kCRk-1) f (COR) where f(COR) = (CORk ? 1)/(COR ? 1). Applying L’Hospital’s rule, with primes 247 esignating preeminence with respect to COR, to the limit of f(COR) as COR
1, yields lim f(COR) = lim (CORk ? 1)’/ Lim (COR? 1)’ = lim kCORk ? 1 = k COR
1 COR
1 COR
1 and limDiesel = 1 ? 1 /CRk ? 1 COR
1 = Otto Thus the limit of the Diesel-cycle efficiency as COR approaches 1 is the Otto cycle efficiency. Hence Equation (6. 18) shows that the efficiency of the Diesel cycle must be less than or equal to the Otto-cycle efficiency if both engines have the same compression ratio, the same conclusion we reached by examination of the p-V diagram.Suppose, however, that the compression ratios are not the same. Compare the Otto cycle 1? 2’? 3’? 4 with the Diesel cycle 1? 2? 3’? 4 having the same maximum temperature in Figure 6. 10. The Otto cycle has a small area, and therefore less work, than the Diesel cycle, but the same heat rejection. Equation (6. 19) demonstrates that the Otto cycle has a lower thermal efficiency than the Diesel cycle with the same maximum te mperature. The conclusion that must be drawn from the above comparisons is quite clear. As in most comparative engineering studies, the result depends on the ground ules which were take at the start of the study. The Otto cycle is more efficient if the compression ratio is the same or greater than that of the competing Diesel cycle. But knock in spark-ignition (Otto) engines limits their compression ratios to about 12, while Diesel-engine compression ratios may exceed 20. Thus, with these higher compression ratios, the Air Standard Diesel-cycle efficiency can exceed that of the Otto cycle. In practice, Diesel engines tend to have higher efficiencies than SI engines because of higher compression ratios. 6. 0 Diesel-Engine Performance In 1897, five years after Rudolph Diesel’s premier(prenominal) patents and blackjack oak years after Otto’s admittance of the spark-ignition engine, Diesel’s compression-ignition engine was be to develop 13. 1 kilowatts of power with an unusual brake thermal efficiency of 26. 2% (ref. 7). At that time, most steam engines operated at thermal efficiencies below 10 %; and the best gas engines did not perform much better than the steam machines. Diesel claimed (and was widely believed) to have developed his engine from the principles expounded by Carnot.He had developed â€Å"the rational engine. ” Whether his claims were exaggerated or not, Diesel’s acclaim was sanitary deserved. He had developed an engine that operated at unprecedented temperatures and pressures, had proven his concept of ignition of fuel by injection into the compressed high-temperature air, and had overcome the formidable worrys of injecting a medley of fuels in appropriate 248 amounts with the precise timing filld for satisfactory combustion. His is a entrancing story of a brilliant and devote engineer (refs. 7, 8).In the Diesel engine, the high air temperatures and pressures prior to combustion are referable to the compression of air alone rather than an air-fuel mixture. Compression of air alone eliminates the possibility of autiognition during compression and makes high compression ratios possible. However, because of the high pressures and temperatures, Diesel engines must be designed to be structurally more rugged. Therefore, they tend to be heavier than SI engines with the same brake power. The energy release process in the Diesel engine is controlled by the rate of injection of fuel.After a brief ignition lag, the first fuel injected into the combustion chamber autoignites and the resulting high gas temperature sustains the combustion of the remnant of the fuel stream as it enters the combustion chamber. Thus it is evident that the favorable fuel characteristic of high autoignition temperature for an SI engine is an unfavorable characteristic for a Diesel engine. In the Diesel engine, a low autoignition temperature and a short ignition delay are desirable. Knock is possible in the Diese l engine, but it is due to an all different cause than knock in a spark-ignition engine.If fuel is ignited and burn down as rapidly as it is injected, then smooth, knock-free combustion occurs. If, on the other hand, fuel accumulates in the cylinder before ignition due to a long ignition lag, an explosion or detonation occurs, producing a loud Diesel knock. The cetane number is the parameter that identifies the ignition lag characteristic of a fuel. The cetane number, like the octane number, is determined by testing in a CFR engine. The ignition lag of the test fuel is compared with that of a mixture of n-cetane, C16H34, and heptamethylnonane, HMN (ref. 0). Cetane, which has good ignition qualities, is assigned a value of 100; and HMN, which has poor knock behavior, a value of 15. The cetane number is then given by the sum of the lot of n-cetane and 0. 15 times the percentage of HMN in the knock-comparison mixture. A cetane number of 40 is the tokenish allowed for a Diesel fuel . 6. 11 Superchargers and Turbochargers The importance of the volumetric efficiency, representing the efficiency of induction of the air-fuel mixture into the reciprocating-engine cylinders, was discussed earlier.Clearly, the more mixture mass in the displacement volume, the more chemical energy can be released and the more power will be delivered from that volume. During the Second realism War, the mechanical supercharger was sometimes used with SI aircraft engines to increase the power and operational roof of American airplanes. Today supercharging is used with both Diesel engines and SI engines. The supercharger is a compressor that supplies air to the cylinder at high pressure so that the as density in the cylinder at the start of compression is hale above the free-air density. The piston exhaust gases are allowed to expand freely to the atmosphere through the exhaust manifold and tailpipe. The supercharger is usually driven by a belt or gear train from the engine crank shaft . 249 Figure 6. 11 shows a modification of the theoretical Otto cycle to jibe mechanical supercharging. The supercharger supplies air to the engine cyclinders at pressure p7 in the intake process 7
1. The processes 4
5
6 purge most of the combustion gas from the cylinder.The most striking change in the cycle is that the induction-exhaust loop is now traversed counterclockwise, indicating that the cylinder is delivering net work during these processes as well as during the compressionexpansion loop. It should be remembered, however, that part of the cycle indicated power must be used to drive the external supercharger. The turbosupercharger or turbocharger, for short, is a supercharger driven by a turbine using the exhaust gas of the reciprocating engine, as shown schematically in Figure 6. 12. A cutaway drawing view of a turbocharger is shown in Figure 6. 3(a). Figure 6. 13(b) presents a diagram for the turbocharger. compendious turbochargers commonly increase the brake pow er of an engine by 30% or more, as shown in Figure 6. 8, where the performance of an engine with and without turbocharging is compared. There, a substantial increase in peak torque and flattening of the torque-speed curve due to turbocharging is evident. For a supercharged engine, the brake power, BP, is the indicated power (as in Figure 6. 11) less the engine friction power and the supercharger shaft power: BP = DISP  IMEP  N ? Pm ?FP [ft-lbf /min | kJ/s] (6. 15) 250 where Pm is the supercharger-shaft mechanical power supplied by the engine (0 for a turbocharger). The IMEP includes the positivistic work contribution of the exhaust loop. The exhaust back pressure of the reciprocating engine is higher with a turbocharger than for a naturally aspirated or mechanically supercharged engine because of the drop in exhaust gas pressure through the turbine. The engine brake power increases principally because of a higher IMEP due to the added mass of fuel and air in the cylinder during combustion.Intercooling between the compressor and the intake manifold may be used to further increase the cylinder charge density. Turbocharging may increase engine efficiency, but its primary eudaemonia is a substantial increase in brake power. In a turbocharged engine, a wastegate may be required to ringway engine exhaust gas around the turbine at high engine speeds. This becomes necessary when the compressor raises the intake manifold pressure to excessively high levels, causing engine knock or threatening component damage. cardinal to forty percent of the exhaust flow may be bypassed around the turbine at maximum speed and load (ref. ). 251 252 6. 12 The Automobile Engine and Air contamination Since the Second World War, concern for environmental pollution has grown from acceptance of the post quo to recognition and militance of national and planetary scope. Among other sources, causes of the well-known Los Angeles smog problem were identified as hydrocarbons (HC) and ox ides of nitrogen (nighttime) in exhaust runs from motor vehicle reciprocating engines. As a result, national and California automobile air pollution limits for automobiles have been realized and toughened.Prior to the dust Air Act of 1990, the U. S. national exhaust-gas emissions standards limited unburned hydrocarbons, carbon monoxide, and oxides of nitrogen to 0. 41, 3. 4, and 1. 0 g/mile, respectively. According to reference 12, today it takes 25 autos to emit as much CO and unburned hydrocarbons and 4 to emit as much dark as a single car in 1960. The reference anticipated that, led by quick California law and other factors, hereafter engine designs should be targeted toward satisfying a tailpipe standard of 0. 5, 3. 4, 0. 4 g/mile. Indeed, the 1990 unaccented Air Act (refs. 15,16) specified these limits for the first 50,000 miles or five years of operation for all passenger cars manufactured after 1995. In addition to the regulations on aerosolized emissions, the prett y Air Act of 1990 adopted the California standard for particulate function of 0. 08 g/mile for passenger cars. The standards on particulates are particularly difficult for the Diesel engine, because of its of soot-producing tendency.The automobile air pollution problem arises in part because the reactions in the exhaust system are not in chemical equilibrium as the gas temperature drops. Oxides of nitrogen, once formed in the cylinder at high temperature, do not return to equilibrium submergences of nitrogen and oxygen in the cooling exhaust products. Likewise, CO formed with rich mixtures or by dissociation of CO2 in the cylinder at high temperature does not respond rapidly to an infusion of air as its temperature drops in the exhaust system. Their concentrations may be thought of as constant or frozen.Unburned hydrocarbons are produced not only by rich combustion but also by unburned mixture lurking in crevices (such as between piston and cylinder above the top piston ring), by lubricating oil on cylinder walls and the cylinder head that absorbs and desorbs hydrocarbons before and after combustion, and by fugacious operating conditions. Starting in 1963, positive crankcase ventilation was used in all new cars to duct fuel-rich crankcase gas previously vented to the atmosphere back into the engine intake system. Later in the ? 0s, different fixes were adopted to comply with regulation of tailpipe unburned hydrocarbons and CO, including sinister compression ratios. In 1973, NOx became federally regulated, and exhaust gas recirculation (EGR) was employed to reduce NOx formation through cut combustion temperatures. At the same time, HC and CO standards were reduced further, confidential information to the use of the oxidizing catalytic convertor. Introduction of air pumped into the tailpipe provided additional oxygen to assist in completion of the oxidation reactions.In 1981, a reducing catalytic convertor came into use to reduce NOx further. This dev ice does not perform well in an oxidizing atmosphere. As a result, two-stage catalytic converters were applied, with the first stage reducing NOx in a near-stoichiometric mixture and the 253 second oxidizing the combustibles remaining in the exhaust with the help of air introduced between the stages. This fresh air does not the increase NOx significantly, because of the comparatively low temperature of the exhaust.The three-way catalytic converter using several exotic metal catalysts to reduce all three of the vaporific pollutants was also introduced. The use of catalytic converters to deal with all three pollutants brought about significant simultaneous reductions in the three major gaseous pollutants from automobiles. This allowed fuel-economy-reducing modifications that had been introduced earlier to satisfy emission reduction demands to be eliminated or relaxed, leading to further gains in fuel economy.Catalytic converters, however, require precise control of exhaust gas oxyg en to near-stoichiometric mixtures. The on-board computer has made possible control of mixture ratio and spark timing in response to shun outputs of intake manifold pressure, exhaust gas oxygen, engine speed, air flow, and incipient knock. The oxygen, or lambda, censor located in the exhaust pipe upstream of the three-way converter or between the two-stage converters is very sensitive to transition from rich to lean exhaust and allows close computer control of the mixture ratio to ensure proper operation of the catalytic converter.Computer control of carburetors or fuel injection as well as other engine functions has allowed simultaneous improvement in fuel economy and emissions in recent years. Thus, while emissions have been drastically reduced since 1974, according to reference 11 the EPA composite fuel economy of the average U. S. passenger car has nearly duple; although this improvement has not come from the engine alone. Despite the hard-won gains in emissions control and fu el economy, further progress may be expected. EXAMPLE 6. 5 The 1990 NOx emissions standard is 0. grams per mile. For an automobile burning stoichiometric octane with a fuel mileage of 30 mpg, what is the maximum tailpipe concentration of NOx in move per million? Assume that NOx is represented by NO2 and that the fuel density is 692 kilograms per cubiform meter. Solution For the stoichiometric combustion of octane, C8H18, the air-fuel ratio is 15. 05 and the molecular weight of combustion products is 28. 6. The consumption of octane is mf = (692)(1000)(3. 79? 10-3)/ 30 = 87. 4 g/mile [Note: (kg/m3)(g/kg)(m3/gal)/(mile/gal) = g/mile. The concentration of NOx is the ratio of the number of moles of NOx to moles of combustion gas products: mole dark /mole cg = (mNOx /mf)(mf / mcg)(Mcg /MNOx) = (0. 4/87. 4)(28. 6/46)/ (15. 05 + 1) = 0. 0001773 254 or 177. 3 parts per million (ppm). _____________________________________________________________________ Bibliography and References 1. Heyw ood, fanny B. , sexual Combustion Engine Fundamentals. raw(a) York: McGraw-Hill, 1988. 2. Ferguson, Colin R. , Internal Combustion Engines. New York: Wiley, 1986. 3. Adler, U. , et al. , automotive Handbook, 2nd ed. Warrendale, Pa. rules of order of self-propelled Engineers. , 1986. 4. Lichty, Lester C. , Internal Combustion Engines. New York: McGraw Hill, 1951. 5. Crouse, William H. , self-propelled Engine Design. New York: McGraw-Hill, 1970. 6. Obert, Edward, Internal Combustion Engines, depth psychology and Practice. Scranton, Pa. : International Textbook Co. , 1944. 7. Grosser, Morton, Diesel: The bit and the Engine. New York: Atheneum, 1978. 8. Nitske, W. Robert, and Wilson, Charles Morrow, Rudolph Diesel: Pioneer of the days of Power. Norman, Okla. : University of Oklahoma Press, 1965. 9. Demmler, Albert W. Jr. , et al. , ? 989 Technical Highlights of Big-three U. S. Manufacturers,? Automotive Engineering. Vol. 96, No. 10, October 1988, p. 81. 10. Anon. , ? Ignition c hoice of Diesel Fuels by the Cetane Method,? ASTM D 613-84, 1985 annual Book of ASTM Standards, Section 5. 11. Amann, Charles A. , ? The Automotive Spark Ignition Engine-A Historical Perspective,? American order of Mechanical Engineers, ICE-Vol. 8, Book No. 100294, 1989. 12. Amann, Charles A. , ? The Automotive Spark-Ignition Engine-A Future Perspective,? Society of Automotive Engineers typography 891666, 1989. 13. Amann, Charles A. , ?The Passenger Car and the Greenhouse Effect,? Society of Automotive Engineers Paper, 1990. 14. Taylor, Charles Fayette, The Internal Combustion Engine in Theory and Practice, 2nd ed. , revised. Cambridge, Mass. : MIT Press, 1985. 255 15. world Law 101-549, ? An Act to Amend the Clean Air Act to Provide for growth and Maintenance of Health, Protection, National Air grapheme Standards, and Other Purposes,? November 15, 1990. 16. Anon. , ? Provisions? Clean Air Amendments,? Congressional Quarterly, November 24, 1990. EXERCISES 6. 1 while dimension less piston position against crank angle for S/2L = 0. , 0. 4, 0. 3, and 0. 2. 6. 2* Obtain expressions for the piston velocity and acceleration as a function of the crank angle, constant angular velocity, and S/2L ratio. Use a spreadsheet to think and plot velocity and acceleration against crank angle for S/2L = 0. 5, 0. 4, 0. 3, and 0. 2. 6. 3 qualify the equation for the piston motion for a scotch yoke mechanism in terms of crank angle. Obtain an equation for the piston velocity for a crank that turns with a given angular velocity,
. 6. 4 Derive an equation for the Otto-engine net work by integration of pdV for the Air Standard cycle.Compare with Equation (6. 6). 6. 5* Use a spreadsheet to calculate and plot cycle efficiency as a function of compression ratio for the Diesel cycle for cutoff ratios of 1, 2, and 3. Indentify the Otto-cycle efficiency on the plot. Explain and show graphically from the plot how a Diesel engine can be more efficient than an Otto engine. 6. 6 A sing le-cylinder Air Standard Otto engine has a compression ratio of 8. 5 and a peak temperature of 3500°F at ambient conditions of 80°F and one atmosphere. Determine the cycle efficiency, maximum cylinder pressure, and mean effective pressure. 6. A six-cylinder engine with a compression ratio of 11 runs at 2800 rpm at 80°F and 14. 7 psia. Each cylinder has a bore and stroke of three inches and a volumetric efficiency of 0. 82. Assume an Air Standard, four-stroke Otto cycle _______________________ * Exercise numbers with an asterisk indic\r\n'