advanced engine technology heinz heisler pdf free download

advanced engine technology heinz heisler pdf free download

Aeronautical Engineering. Automobile Engineering. Civil Engineering. Computer Engineering. Electrical engineering. Environmental Engineering. Industrial Engineering.

The primary chain provides a parallel chain drive conveying motion from the crankshaft sprocket of, say, 21 teeth to the larger of the two intermediate sprockets of, say. The scoondary chain then transfers the chain drive from the smaller intermediate sprocket wheel of, say, 20 teeth upwards and outwards to the cam shat sprockets having, say. Primary chain slackness is supported by a slipper-head chain tensioner on the outside, just above the crankshaft, whereas a vibration damper Strip is positioned on the inside of the chain just below the intermediate sprockets, A third chain is used to provide a speed step-down and drive for the lubrication pump.

Twin overhead camshaft and auxiliary equipment two-stage chain drive Fig. A triple sprocket triangular-chain drive forms the primary stage, the crankshaft sprocket providing the input drive to the larger intermediate double sprocket wheel and the auxiliary shaft sprocket.

No idler jockey sprocket is used as in Fig. The deep mounted idler jockey sprocket provides a large angle of wrap for both camshaft sprocket wheels. Chain slack control is achieved by the long curved strip chain tensioner and chain vibration is res- trained by the normal vibration damper guide strips.

A separate oil pump chain drive can be Fig. Provided twin camshafts are not essential, the single-chain drive weighs far less, and costs are reduced, compared with the twin overhead cam: shaft wo-stage arrangement shown in Fig. Vee 8 and 12 twin overhead camshaft two-stage chain drive ig. A fourth simple short-chain drive rotates the oil pump, a large oil-pump sprocket being used to reduce the oil-pump speed so that the power consumed by the pump is minimized.

Thus, the chain moving from the crankshaft to the camshaft. However, generally, the deceleration period is only of a short duration and is therefore not a major problem. Consequently, chain tensionets are incorpo- ted on the normally slack side of the chain to take up any chain slack, thereby restricting chain whiplash Fig, 2,25 b.

In severe operating con- ditions guide strips may be provided on the drive side of the chain span between the sprockets 2. A cylindrical slider-block with rectangular elongated central slot fits over the flanged mounting Fig.

A rectangular twin pawl housing is screwed onto the end of the flanged mounting, whereas pair of segmental ratchet plates are screwed on either side of the pawl housing against the slider-block.

This, there fore, restricts both the slider-block and sprocket wheel end-float respectively. When assembled, the ratchet-pawls are forced outwards by the central bias-spring until they both engage the ratcher-plate teeth. Inservice, as the chain stretches, the compress ion-spring situated between the flanged mounting and the slider-block pushes the slider block and sprocket wheel assembly further towards the chain to compensate for the increased amount of slackness in the chain.

As the slider-block moves progressively over, both toothed pawls ride over the segmental ratchet-plate teeth until they align and drop into place with the next set of meshing teeth. Consequently, the tensioner sprocket- 2 Timing chain cesponse-under operating conditions wheel assembly is permitted to move in one direction towards the chain but is prevented from retracting rearwards under chain backlash condi- tions 2. Ratchet strip and block automatic chain tensioner Bigs 2.

Thus, as the two arms move further apart with increased chain slackness, the block slides further betwcen the inclined arm and over the ratchet strip to engage the next adjacent teeth. Located inside the plunger is a ratchet sleeve which has a slot spiralling from top to bottom with a series of stepped semi-circular notches formed along one edge of the spiral a hetix. Whip-back of the is prevented by the limit-peg--protruding inside the plunger near its mouth—being pushed back with the plunger and slipper head unti it aligns and sits in one of the ratchet-stepped semi-circular notches Fig.

By these means the plunger and slipper head auto- matically move outwards, whereas the ratchet sleeve revolves as the chain and sprocket teeth wear, Because of this, the backlash retraction of the plunger remains constant as the plunger pro- gressively extends from its eylinder housing to compensate for the increasing chain slackness.

Before the plunger is inserted into its cylinder housing during assembly. After the tensioner assembly has been bolted to the engine block. Lubrication oil from the engine is supplied to the automatic ratchet mechanism via a drilling in the backplate, and it passes out through hole in the slipper head. The body of oil tends to damp down any violent to and fro movement of the plunger and also lubricates the slipper head and chain. Back pate Fig. Ilustration of primary and secondary motion 3.

Primary and secondary piston movement 3. Next, draw the line-of-stroke through the centre of the circle, divide one half of the crankpin circle into a number of equal parts at, say. Set a compass to the sealed connecting-rod length and mark off the small-end gudgeon pin positions along the ex- tended line-of-stroke for corresponding crankpin big-end angular dispositions.

Draw perpendicular Tines to the line-of-stroke for each incremental position of the smallend gudgeon-pin, also draw a second set of lines parallel to the line-of-stroke passing through the crankpin points until they intersect the corresponding perpendicular lines for each gudgeon pin position. Draw projected lines through each crankpin centre interacting the perpendicular centre-line and mark these points 0,6 1.

Set the compass to the scaled connecting-rod length and mark off the piston gudgeon-pin posi- tion along the line-of-stroke for each intersected point on the perpendicular centre-line Secondary piston movement may be defined as the piston travel caused by the rotating crankpin's, projected positions on the perpendicular centre- line. The secondary movement velocity graph can then be produced by drawing lines perpendicular to the line-of-stroke through each.

Xp —X, When the piston reaches the end of the out- ward stroke and commences its return inward stroke from BDC to TDC, the relative displace- ments of the primary and secondary movernents are similar to those on the outward-going stroke. Thus, when moving from BDC to mid-crankpin position the secondary movement reduces a por- tion of the primary movement and, with further movement from mid-crankpin position to TDC, 93 the secondary movement then, in effect, extends the primary movement.

Set the compass t0 the scaled connecting-rod length and mark off the small-end gudgeon-pin positions along the line-of stroke for corresponding crank- pin big-end angular dispositions. From TDC extend back these lines until they intersect the big-end circle centre perpendicular line and mark off these points, Draw petpendicu- Jars along the line-of-stroke for each incremental gudgeon-pin position and then draw lines parallel to the line-of-stroke passing through the pro- jected erankpin intersecting points on the perpen- dicular erankpin circle line.

The resulting velocity graph is then produced by drawing a smooth curve through the perpendicular and parallel intersecting points, With the piston stroke as the base the magnitude of the piston velocity at any position of the gudgeon-pin is presented by the perpendicular distance between the line-of-stroke and the curve.

The increased distance travelled by the piston for the first half of the crankpin angular move- ment compared with that moved in the second half Fig, 3.

Decel Doce! Consequently, there is a relative imbalance of acceleration and decelera- tion when the piston changes its direction of motion at either dead-centres. Let x be the displacement of the piston from the inner dead-centre position Fig. However, its peak magnitude is only between a quarter to one-third of the primary acceleration depending upon the connccting-rod length to crank-throw ratio. Algebraically combining both the positive and negative accelerations of the primary and secondary piston movements over one com- plete revolution produces an increase in piston acceleration as it moves out from TDC—this being shown by the height and steepness of the resultant curve for the first crank-angle quadrant In contrast, the resultant deceleration of the piston as BDC is approached is reduced, this being shown by the almost flat plateau at the base of the resultant acceleration curve.

Consequently, the crankshaft will move through four instead of two equal peak out-of-balance positions every revolution—that is, at TDC and BDC in the vertical plane caused by the reciprocating masses and at? Horizontal opposed twin changes its direction, at mid-stroke. Method of solving the state of balance for primary and secondary forces and couples for various crankshaft and cylinder configurations Figs 3.

Draw dimension lines parallel to the crankshaft length and insert distances xy x2. Primary and secondary force vectors fare treated as acting radially outwards from crank centres. However, the secondary force on piston No. Consequently, the offset between the two pistons causes the primary forces which are parallel but point in opposite directions to each other to produce a moment of a couple.

In contrast, both secondary forces act in the same direction and therefore they produce a vertical shake. With the crankpin in the TDC posi- tion Fig. However, there will be a secondary inertia force on piston No. Correspondingly, piston No. With this simple crankshaft arrange- ment both primary and secondary inertia forces are balanced as shown by the two force polygons Fig. In contrast. The out-of- balance couples are por- trayed by the open primary Cp and secondary C, moment polygons as shown Fig.

The couples produced as the crankshaft rotates will be wz of varying magnitudes, as explained in the fol lowin; When piston No. I and No. The lower pistons. However, due to all the pistons being in different planes the downward force of the first two pistons is opposed by the upward force of the rear piston, the result being that an anticlockwise pitch couple is produced which tends to tit the engine downward at the front and upward at the rear Fig.

Accordingly, the maximum down- ward inertia force of piston No. I's TDC position brings piston No. In this position, the upward force of piston No. The resulting offset upward force components and the max- imum downward piston force along the crank: shaft length now produces clockwise pitch couple which tends to tit the front end up and the rear end down Fig.

Thus, the cycle of events is repeating Fig, 3. In contrast, the secondary forces imposed on pistons Nos 3 and 4 act in opposition to their respective primary force since both of these pistons are at BDC and therefore their sense of direction must be towards their respec tive TDC positions.

Thus, it can be seen from the force and moment polygon diagrams that only the primary forces are balanced and that the primary couples, secondary forces and secondary couples are unbalanced Fig. Likewise, piston No. I's TDC position in the left-hand cylinder- bank. In this crankshaft position the pistons in the leftchand cylinder- bank generate a small horizontal clockwise couple while the pistons in the right-hand eyfinder-bank will generate a large horizontal anticlockwise couple.

The unbalanced primary couples can- not be simply balanced by attaching counter- Primary force moment polygon Polygon Fig. Draw the primary crank PC , which is the end view of the crank- shaft, showing cach crank-throw arm with its anguiar displacement 8 from the line-of-stroke.

Finally, draw the secondary imaginary crank SC. Primary and secondary force polygon Fig. From the end of the first line draw to scale a parallel line to the crank arm 2 representing the piston primary force 2 and insert an arrow pointing again in an outward direction from the centre of the primary crank. From the end of the second line draw to scale a parallel line to the crank arm 3 representing the Piston.

Repeat this procedure for primary moment vectors 3. Summary of primary and secondary force and couple balance Referring to Fig. It can also be seen that the secondary moment polygon is closed so that there is no unbalanced secondary couple. When piston No. Thus, due to the staggering of the pistons over the crankshaft span, a large anti- clockwise pitching couple is produced. Rotating the crankshaft further, to.

I's TDC position, brings piston No. This combination of vertical in- ertia forces which are equally spaced and are of equal magnitude cancel out any likely pitch couple, Hence, there is no pitch couple while the crankshaft remains in this position. However, when two of these three-cylinder crankshafts are joined together they may be tre- ated as the mirror image of each other with the first and last , adjacent middle and in between front and rear crankpins paired Fig.

Therefore, the left- and right-hand cylinder banks are staggered to accommodate and align the connecting-rods with their respective crank Primary Primary force moment polygon polygon pins, the left-hand bank being slightly forward relative to the right-hand bank In effect, the left- and right-hand cylinder banks consist of a pair of three-cylinder in-line engines with each set of crank-throws being a mirror image of the other.

As a result, the unbalanced couples produced by cach cylinder bank. The left-bank is numbered from the front to the rear 1, 2 and 3 and the right-hand bank also follows this pattern with numbers 4, 5 and 6. Thus, the left-bank cylinders align with the fist, third and fifth erankpin whereas the right- hand bank eylinders align with the second, fourth and sixth crankpin, counting from the front of the crankshaft. Conversely, however, there is a large primary couple created as shown by the open primary moment polygon Fig.

Conse- quently, there will be angular periods when the direction and magnitude of these couples do not cancel each other out, this then causes periodic external couples to be generated. With counterweights forming part of the fly wheel and rear three extended crank webs, and Giagonally opposed counterweights imposed on the front pulley and first three extended crank webs, primary couples can be neutralized Fig.

Thus, both primary forces and couples are balanced Fig. This is shown by the closed polygon force and moment diagrams. However, the secondary force is unbalanced as the force diagram does not close, whereas that of the secondary moment polygon closes, indicating thet the secondary couples are neutralized 3. These couples which are at right-angles to each other therefore produce a relative moderate unbalanced resultant couple. I's TDC position now generates a large anticlockwise couple in the right-hand cylinder bank which is partially coun- teracted by a small clockwise couple coming from the right-hand cylinder bank; consequently, there will again be an anticlockwise resultant couple, but of a smaller magnitude.

However, all the primary and secondary shake forces are balanced and there is no secon dary rocking couple shown by the closed force and moment polygon diagrams, Fig. Oey Fig. Thus, when any of the pistons are at TDC or BDC their primary inertia force will be exactly balanced by the counterweight mass, while the adjacent piston on the opposite cylinder-bank will be in its mid-stroke position where there is no primary piston inertia force.

However, the secondary inertia force of each piston at mid-stroke is negative see Fig. Conversely, when the piston is at either TDC or BDC the secondary force is positive and at a maximum and points upwards along its line-of-stroke.

Therefore, it can be seen that, with the use of counterweights, all primary and secondary forces and couples are eliminated Fig. Secondary Secondary force moment polygon polygon 12 3.

Pairs of traversely located pistons are attached to each crankpin via their connecting-rods. Hence, the engine is perfectly balanced for primary and secondary inertia forces and couples Fig. This erankshaft and cylinder configuration therefore provides the very best smoothness and acceleration response of all the engines examined 3. These gears are then made to rotate av the erankshaft speed by meshing the left-hand countershaft to an input gear, which is of the same size as the countershaft gears and, which is itself, mounted on the crankshaft.

Correspondingly, the primary negative inertia force which now has a downward. BDC direction will be neutralized by the upward direction of the centrifugal forces produced by both counterweight. The countershaft is driven at crankshaft speed in the reverse direction by a pair of meshing identical ear wheels attached to the front end of the crank and bakance-shaft. Weights are attached in the form of extended crankwebs on both sides split weights of crankpins Nos 1 and 3.

Likewise, a pair of diametrically opposed weights are attached to the front and rear end of the balance countershaft. Therefore, the upward these centrifugal forces cancel-out primary maximum force of piston No.

Rotating the crankshaft 60 from crank No. In this position the primary whereas the downward direction of pistons Nos maximum downward force of piston No. The position brings piston No. In this position the primary max- imum downward force of piston No. However, the front and rear balance weights now act in the opposite sense to cancel out the primary couple and thus eliminate any vertical pitching Rotating the crankshaft " from crank No.

Correspondingly, the balance weights of the crankshaft and countershaft will be in the horizontal plane where their centrifugal forces cancel each other. Consequently, the centrifugal forces of these weights will now oppose the downward dip at the rear and the upward lift at the front, thus neutralizing the vertical primary couple 3.

Thus, the dual stiffness of the crankshaft system produces a second frequency of vibration, which results in a frequency below and above the original frequency when no rubber vibration damper is attached.

Figure 3. Likewise, heat reduces both the stiff ness and the damping ability. Rubber also has the highest internal friction hysteresis of all the engineering materials. To achieve these fundamental requirements the designer and development engineer has to be aware of the factors that contribute towards in: ducing the charge to enter the cylinder, to mix intimately, to burn both rapidly and smoothly and, 10 expel the burnt gases.

Induction swirl is pro- duced by positioning the induction port passage 12 to one side of the cylinder axis so that the flow discharges into the cylinder tangentially Fig. Directed straight port ig, 4.

It then discharges into the cylinder tangentially towards the cylinder Wall where itis deflected sideways and down- wards in a spiral or whirling motion Deflector wall port Fig. It then discharges into the evlinder with a predeter- mined downward spiralling motion about the eylinder axis. Terminology of airfmixture movement inside the cylinder axis.

The closing of a portion of the valve seat periphery restricts the flow discharge and there- fore reduces the possible volumetric efficiency of the cylinder but with the benefits of a large amount of swirl Helical port Figs 4.

The charge flow in the induction port is guided by the passage walls, which make it spiral around and. The intensity of swirl is influenced by the steepness of the port helix and the mean diameter Of the spiral low path about the valve axis.

Helical ports usually provide higher flow dis- charges for equivalent levels of swirl compared with directed ports because the whole periphery of the valve opening area can be fully utilized and. Helical ports are less sensitive to their position relative to the cylinder axis since the swirl gener- ated depends mainly on the port gcometry above the valve and not how it enters the cylinder.

The nose of the steep chamber wall partially separates the inlet and exhaust valves so that the incoming charge is forced not only to move downwards but also 10 rotate around the wall of the cylinder.

Thus, the resultant downward and circular movement of the mixture generates an expanding and then a con- tacting spiral swirl about the cylinder axis during both the induction and the compression strokes, respectively. In contrast to induction, during the exhaust stroke, the burnt products of combustion are gently directed by the vertical circular chamber wall and the curved roof of the chamber so that 0 Defector wall port 8 Hotical pot the least resistance is experienced during the expulsion of the exhaust gases.

Directed induction switt space situated either in the piston or eylinder head. Quench area Fig. Consequently, there will be a large amount of heat transferred from this thin lamina of hot charge through the metal walls, The result is a rapid cooling or quenching effect, by these parallel surfaces.

The quench area is defined as the percentage of opposing flat area relative to the piston crown Turbulence Fig, 4. These vortices, which are carried along with the flow stream, represent small irregular breakaways that take on a concentric spiral motion Fig. The stroke-to-bore ratio for va rious engines can range from 0.

However, this advantage 6 Uncer square engine Fig. A furth- er consideration is that peak cylinder pressure tends to decrease as the stroke-to-bore ratio becomes more undersquare, 4. The exit between the mating seats then results in the formation of a free jet. At low exhaust valve lift Fig. Consequently, the sudden expansion of the exhaust gas dissipates the kinetic energy of the gas with only the minimum being converted to prssure energy downstream of the valve port seat.

However, as the throttle valve is steadily closed there is a corresponding inctease in the cylinder and manifold depression during the induction stroke. Thermal efficiency 5 6 engine load Fig. With the petrol engine, a mixture of air and fuel is drawn into the cylinder during the outward moving induction stroke.

On its return inward moving stroke, the mixture is compressed to Something like one-ninth to one-eleventh of the tunswept volume, and just before the piston reaches the end of its compression stroke a spark isused to ignite the combustible mixture In contrast.

Just before the piston reaches the end of the com- pression stroke an accurately metered quantity of fuel is injected into the cylinder at pressures of bar or more. The finely atomized fuel spray mixes with the hot air causing it to ignite due to the heat and burn rapidly. To minimize the mixing of the fuel particles, the air is given movement so that it sweeps across the finely atomized and penetrating spray.

However, at high engine speeds, with such short injection and mixing times, the chemically correct amount of air charge will not be able to seek out, intimately contact, mix and burn all the fuel particles; therefore, very high unacceptable levels of soot and black smoke would be expelled with the exhaust gases. Consequently, diese!

This means that for the same power output. Heat is transferred from the metal surfaces to the entrap- ped homogencous mixture which is uniformly distributed throughout the cylinder. However, the ignition tempera ture does not rely on the cylinder compression pressure alone but more so on the heat transfer in the cylinder, which itself increases with the dens- ity of the air fuel mixture. This contrasts with diesel compression ignition engines.

Thus, with the diesel combustion process. The reduced rates of burning with these extreme mixture strengths Fig. Compression ratio Figs 4. Correspondingly, the maximum cylin- der pressure inereases from 32 bar to 82 bar and the brake mean effective pressure generated also increases from about 9.

AS would be expected, raising the cylinder temperature reduces the ignition delay period for one set engine speed Fig. Thus, for an engine running in its mid-speed range, the ignition timing would be reduced from It therefore raises both the engine ther- mal efficiency and the developed power. In theary. One further effect of raising the engine com- pression ratio is that it reduces the eylinder clear ance volume.

This results in a lowering of the der volumetric efficiency since less. L Fig. The combustion process may be considered to lake place in three phases or periods: I the delay period; 2 the rapid pressure rise period; 3 the after burning period, 4. Delay period ignition and early flame development 0 Figs 4. This period tends to be very nearly constant in time The duration of this period is dependent upon.

Thus, the time interval of burning from the point of ignition 0 peak pressure rapid pressure rise period reduces, while its duration, expressed in degrees of crank-angle movement, remains roughly constant.

Thus, by automatically advancing or retarding the actual firing point with increasing or decreasing speed the optimum engine torque will be main- tained.

Generally, the detonation pressure oscillations when superimposed on the normal pattern of combustion pressure rise and fall will show a slightly higher peak eyclic eylinder pressure. During the combustion process, the atomized homogeneous fuel-air mixture is ignited by the passane of the spark between the spark-plux electrodes.

If there is no movement of the fresh mixture in the cylinder the nucleus of flame initiated from the spark would form an unbroken burning front travelling progressively outwards to the furthest point in the combustion chamber at a very low velocity However, the flame front is normally subjected to turbulence which distorts the smooth flame front into a ragged and broken one Fig. These shockwaves reflect against the com: bustion chamber walls and. Therefore, the majority of the charge will barn in a controllable manner and only the very last of the unburnt charge may actually reach the spontaneous condition which causes dctona- tion.

Thus, the earlier in the combustion process the detonation commences, the more unburnt end-mixture will be available 0 intensify the detonation. Accord ingly, if this is excessive it will destroy the gudgcon-pin boundary lubrication and hence cause abnormal wear to the small-end joint. Factors which promote detonation 1 High compresstion ratios. However, the combined effect of shorter com- pression time and reduced volumetric efficiency, as opposed to increased cylinder temperature favours reducing the detonating tendency in the oylinder with rising engine speed 4.

The accumulated effects of an extended com- bustion time and rising peak cylinder pressure and temperature, cause the sell-ig ture to creep further and further ahead of TDC. Under these conditions, when the engine is driven hard. Preignition is initiated by some overheated projecting part such as the sparking plug elec- trodes. The resulting in- creased heat flow through the walls.

Preignition is not responsible for abnormally high cylinder pressure, but there can be a slight pressure rise above the normal due to the ignition point and. If preignition occurs at the same time as the imed sparking plug fires, combustion will appear as normal, Therefore, if the ignition is switehed- off, the engine would continue to operate at the same speed as if it were controlled by the conven: tional timed spark, provided the self-ignition temperature continues to occur at the same point.

After a time with the throttle closed. A weak, slow-burning, mixture or retarded ignition setting can be responsible for the overheated hot-spot and thus for the occurrence of post- ignition. The octane number is based on two hydrocarbons which have very different knock properties enabl- ing them to define the ends of the scale: iso- octane CyHyq which has a very high resistance to knock and therefore is given an octane number of , and normal heptane CyHyq which is very prone to knock and is therefore given a zero value, Blends of these reference fuels define the knock resistance of intermediate octane numbers.

Variable compression ratio single-cylinder rosearch engine mixture strength to give maximum knock re- sponse. Various reference fuel blends are then tested in the same way until the compression ratio of one of these reference blends exactly matches the commercial fuel in terms of si jock intensity. This fuel is then rated by the octane percentage in the equivalent biend of reference fuel. The relationship between octane number and compression ratio is approximately as shown Fig.

It is considered to be similar to the detonation tenden. It is considered to be similar o the detonation tenden- cy of a fuel when the engine is driven at medium speed in top gear with a wide-open throttle under heavy load. The more severe operating conditions for the Motor method compared with the Research method will therefore predict a lower octane number for the MON than for RON.

There- fore, chassis dynamometer tests which predict road conditions have been devised to determine the fuel's road octane number. It has been found that the road octane number lies roughly between the Research and Motor octane ratings.

The effects of cylinder pressure on detonation and octane number requirements Figs 4. Generally, the octane number fuel require ments for an engine, depend upon the maximum indicated mean effective pressure generated in- side the cylinder. Thus, the highest knock-limited indicated mean effective pressure Fig. A lower octane number fuel can be used by retarding the ignition timing slightly The mixture strength influences the knock- limiting indicated pressure Fig.

However, for light-loads and maximum economy conditions, weak mixtures which burn slowly are used. Effect of air-fuel mixture strength on the knooklimited imn. Consequently, research en- gineers have been striving to reduce the respon- sible pollutants emitted from the exhaust system without sacrificing power and fuel consumption. Pollutants are produced by the incomplete burning of the air-fuel mixture in the combustion chamber.

The major pollutants emitted from the exhaust due to incomplete combustion are: un- burnt hydrocarbons HC , oxides of nitrogen NO, and the highly poisonous carbon monoxide CO. If, however, combustion is complete the only products being expelled from the exhaust would be water vapour, which is harmless, and carbon dioxide, which is an inert gas and, as such, is not directly harmful to humans.

The conflicting characteristics of the three ex hhaust pollutants, the fuel consumption and de- veloped power and how they are influenced by the mixture strength ate illustrated in Figs 4. Con- versely, under part throttle maximum economy conditions with weakening air-fuel mixture ratios, from about Thus, there has always been a major conflict between the high carbon monoxide con- centration with rich maximum power mixtures and the high nitrogen oxides concentration with Jean part-throtile maximum economy mixtures.

This, therefore, shifts the air-fuel ratio to the decreasing portion of the nitrogen oxide curve about half-way down, whereas carbon monoxide is at a minimum and the unburnt hydrocarbons are only just on the increasing upward slope Fig. With further improvements in promoting in duction swirl Fig. Conversely, with conventional combustion chambers the burning time for very lean mixtures, under part throttle, is far too long so that the rapid pressure-rise period occurs when the clear- ance volume is itself expanding fairly quickly.

However, lean-burn combustion chambers are able to promote large amounts of induction swirl Fig. Ratio of tilted triple valves to tilted single valve seat bore Axou An 3x Anas seas ead this amounts to a If the inlet valve seat bore cross-sectional area of the flat upright single inlet valve cylinder- head is taken as unity, then all other valve configurations, which enlarge the total intake valve seat bore areas for the same cylinder bore size, can be considered as an improvement. Twwo-valve hemispherical combustion chamber Fig.

Within the Inditectinjection swiri combustion chamber ilustrating phases of combustion Ricardo Comet cylindrical casing is a spherical chamber with a narrow parallel passage or throat leading to the radial nozzle holes. A transverse bar with a spherical bulge in the middle ball bar is posi- tioned in the lower half of the spherical chamber Whereas a cold start heater plug intersects from the side of the upper half of the chamber wall When the inlet valve opens and the piston moves away from the cylinder-head, air enters the cylinder tangentially so that it rotates in a down- ward direction about the cylinder axis Fig.

Almost immediately. The thrust of combustion projects these direc- tional jet-like flame-fronts towards the cylinder walls and, in doing so, sweeps the burnt gases and soot to one side while exposing the remaining fuel vapour to fresh oxygen Fig.

With the diesel engine. The different points on the consumption loops are listed below Fig. In contrast with the diesel engine, as the load is reduced the single consumption ioop only begins to rise when the b. This shows that the Petrol engine has considerable amounts of carbon monoxide CO formed as the mixture moves from the stoichiometric This therefore causes the flow of incoming charge tolag behind the piston movement.

Consequently, the depression created by the slower moving charge being unable to keep up with the rapidly expanding cylinder space reaches a maximum of about Beyond BDC the late closing of the inlet valve wall permit the charge, due to its inertia, to continue to enter the cylinder against the returning piston which is now on its eompress- ion stroke.

The graph also shows that, in terms of piston linear movement, the cylinder depression appears to commence very early in the induction period and continues almost horizontally right to the outermost piston position at BDC. Only at the point when the closed throttle compression curve interscets the atmospheric line does the effective compression stroke actually commence which, inthis example.

Hence, the true compression ratio based on the effective part of the compression stroke length with closed throttle will be far less. However, a hori zontal draught and, in particular. However, a hori- zontal draught and, in particular. Hence, a larger dia meter riser can be used with a horizontal draught and downdraught manifold, as lower minimum air velocities of something like 10 mis can effec tively be employed. The charge flow velocity through a manifold riser during wide open throttle conditions de- pends on the total number of cylinder displace- ments, drawing from one manifold riser.

The flow of air along a pipe tract from the walls to the centre varies considerably, in fact the air velocity is greatest at the centre of the passage and decreases to almost nothing on the pipe walls. Hence, fuel particles find it easier to remain in suspension if discharged towards the centre of the tract. Manifold branch flow paths 5.

The Fig. The maximum charge velocity in the narrow section of the tract is high and the velocity profile decreases gently, initially from the centre, followed by a more rapid decline as the tract walls are approached. The decrease in flow velocity from the narrow to the wide tract makes it more difficult for the suspended fuel Particles to be supported by the slower moving air stream, therefore they will precipitate and be forced by the viscous drag to accumulate on the tract walls, Flow in converging tract Fig.

The maximum velocity in the wider cross-section intake tract is lower than for the natrow outlet tract, but in the funnel Fig. Flow ina converging tract or tapered transitional region, fuel particles may accumulate on the converging walls before being drawn in spurts into the high-velocity flow narrow part of the tract. Internal wall surface finish Figs 5. This effect becomes more pronounced with increased throttle opening. With rough surface finishes Fig, 5. This process of build-up and breakaway of liquid fuel from the tract walls produces a con- tinuous and effective mixing mechanism for the charge as it moves towards the inlet port Conversely, a smooth surface finish reduces the surface flow resistance keeping the film thickness 24 to a minimum, consequently there will be very little mechanical break-up or mixing of the fuel particles with the air stream until the erratically suspended fuel particles impinge on the under- side of the inlet valve head.

Section shapes Rectangular sections Fig. One benefit of a rectangu- lar section is that it prevents the column of charge swirling as it moves through the tract, thus mini mizing any centrifuge effect which would force the heavier liquid particles to be flung against the tract walls Circular sections ig. They therefore offer the least resistance to charge flow and this gives the highest volumetric efficiency and, correspondingly.

However, the charge col- Fig. As a result, the distribution of mixture strength across the section may be very uneven. A disadvantage of the Fig.

As a result, the tendency for the liquid fuel particles to. This mini: End-sections Fig. Varying rectangular section elbow mizes liquid fuel particles being precipitated onto the inner bend walls and yet it maintains the Towest flow resistance as the charge flows gently through the curved tract passage of the varying wall shape, but which has a constant cross- sectional area.

However, when the induction period commences on the inner cylinder, charge will be drawn into the leading side of the sharp junction and a small portion of the overshoot on the trailing side of the junction will reverse its Vortex filament Fig. The flow resist- ance for the first sharp junction branch is higher than for the second curved tract and, in addition, with the overshoot charge beyond the first branch tract, the end tract provides a slightly higher filling capacity than the first branch.

With the second curved branch tract there will be some vortex filament and liquid film formation at the beginning of the inner bend.

Mixture distribution at the entrance of each branch is fairly even up to part throttle and light load conditions but it becomes sporadic as the throttle opens further with higher engine speeds. Under certain open- throttle conditions there may be some vortex filament formation and a very small amount of liquid fuel clinging to the inner prong walls of the junction, and on the outer walls just beyond the parallel portion of the tract where the prong divides the branches.

This is because the mixture which cannot get through the upper hall-fap gap which in effect forms a sharp edge orifice then flows to the opposite lower half-valve side of the valve where it finds it easier to wedge itself through the valve opening due to the funnel- like entry shape. Thus, if the throttle valve spin- dees positioned at right angles to the tee junction branches Fig.

If the throttle valve spindle is positioned parallel with the tee branches Fig. Consequently, the manifold tracts will be subjected to a high state of depression with the result that the fuel spray discharged from the carburettor will be finely atomized.

The flow of mixture between the lower half-valve edge and riser wall forms a tapered streamlined orifice, whereas the flow gap between the upper half- valve edge and riser wall forms a sharp orifice Consequently, more mixture flows past the lower half-valve side than that of the upper half side, this therefore causes an uneven charge flow veloc- ity across the riser tract.

The effects of throttling the flow through the riser are therefore twofold 1 The partly opened butterfly valve presents a central obstacle to the mixture moving through the riser, the mixture therefore becomes turbu- lent downstream of the throttle valve, 2. The partly opened butterfly valve produces an uneven flow between the two half-streams of mixture with the result that the converging mixture streams take on a swirling motion as they enter the branch tracts, The net result is that the central air stream in the branch tract tends towards a series of vortex.

Full throttle conditions Fig, 5. Under these conditions, the mixture flow velocity will only be moderately high, and consequently the manifold depression will be relatively low. The fuel spray discharged from the carburettor with the low state of depression is, relatively coarse and the air movement is not totally sufficient to support the liquid fuel parti- cles through the riser and branch tracts.

It there- fore follows that some of the fuel particles will be precipitated from the mixture stream onto the manifold tract walls. Under these conditions, the scrubbing action of the air stream moves closer to the tract walls and excites the liquid flim and air into vortex films at the bottom of the riser and downstream of the riser tee junction as well as on the inner bends of the curved branches.

These vortex liquid films grow in size and periodically breakaway from the walls and are then gulped back into the main misture stream, However, these globules of liquid fuel do not necessarily disintegrate into very small particles as the mov- ing column of the mixture in the tracts is drawn into the cylinder. Generally, the mixture distribur tion may be irregular with a wide-open throttle at only medium speeds, but at higher engine speeds the quality of the mixture thoroughness in mix- ing tends to improve 5.

If the cylinders are joined by a single induction mani- fold, then the second eylinder to commence its induction period interferes with the first cylinder induction period and therefore will be responsible for the uneven filling between cylinders, 5. Induction period overlap interference for a four-cylinder engine can be overcome by having dual intake risers where the two outer cylinders are fed by one carburettor barrel and the inner adjacent cylinders are fed by a second carburettor barrel.

However, twin carburetors, where one carburettor feeds number one and two cylinders and the second carburettor supplies cylinders three and four, do not solve induction overlap interference between cylinders. Thus, with a firing order of , consecutive induction periods will be drawn from left and right-hand manifolds in turn. That is, cylinders 1, 2 and 3 draw from the front half manifold whereas cylinders 4, 5 and 6 draw from the rear half-manifold.

Likewise, a vee six-cylinder engine Fig. Vee eight-cylinder engine Fig, 5. This is represented by the higher volumetric efficiency Fig, 5. However, because of the increase in flow resistance with rising engine speed and the grea- ter amount of flow resistance with the longer tracts, the longer outer branches cause the volumetric efficiency in the cylinders they feed to peak earlier than for the inner cylinders.

Slightly later during the induction period this exhaust gas is pulled back into the various cylinders with the fresh incoming mixture. Under part-throttle conditions, the residual ex- haust gases at the end of the exhaust stroke and the beginning of the induction stroke will initially draw exhaust gases into the induction manifold and then back into the cylinder, the actual amount of gas moving to and fro will not always be equal, and this is mainly responsible for the poor mixture distribution during idling and part throttle opening conditions.

Mixture distribution is generally best with a wide-open throttle where the throttle valve offers the least interference with the incoming air or mixture flow stream. However, the walls of the venturi and induction manifold passage walls will tend to be wet under these particular operating conditions.

With reduced throttle opening the butterfly valve obstructs the incoming mixture stream and therefore causes an uneven quantity of mixture flow between both sides of the tilted butterfly valve plate. Therefore, the mixture distribution between induction branch tracts may vary con- siderably Fig. This is caused by the variation in the vacuum existing in the individual branch pipes due to their flow path-length, passage curvature and shape, and how they converge to form the common mani- fold.

The walls of the induction riser and branches tend to be fairly dry when the depression in the manifold is relatively high Figure 5. This is because the mixture moves from the riser into the gallery and then outwards, and some of the mixture is then diverted to the tee junctions of the inner branches. The remain: ing outward moving mixture is then redirected through the end clbows to the outer branches.

The outward accelerating piston 4uickly expands the space between the eylinder- head and piston crown. Thus, the Kinetic energy gener- ated by the fast-moving column of charge is now converted into pressure energy in the blanked-off inlet port. Consequently, the density of the trap- ped charge rises. For a given tract length, the greater its dia- meter the larger wil its surface area ex- posed to the air stream and the higher will be its dow resistance.

The rate of volumetric efficiency increase with rising engine speed for a fixed mm tract length but with different tract diameters, 30, 40 and 50mm, appears to be.

A small-bore tract will have a relatively high flow velocity so that itis well able to maintain a thoroughly mixed charge of air and fuel in fan atomized state as it travels through the tract at low engine speeds Fig. In 27 contrast, a large-bore tract will not be able to hold the liquid particles in suspension as the charge flows through the tract at low engine speed.

Consequently, the heavicr liquid part cles may precipitate onto the walls of the tract. However, the charge velocity and the accompanying fiow resistance become exces- sive towards the maximum engine speed, which therefore results in a rapid decrease in cylinder volumetric efficiency 5. Thus, with a firing order of , the mixture is drawn first from the left-hand side main gallery cylinder number 1 then from the right cylinder number 3 main gallery and right 4 again, then left 2 and left 1 again, it then repeats this sequence L-R-R-IR-RUL etc This pattern of mixture draw from the main gallery of the manifold tends to produce unequal mixture distribution between branches.

Four-branch induction manifolds with a single carburettor Fig. However, the two outer cylinders will have a fonger flow path than the inner ones. The separa- tion of port branches as opposed to the siamese inlet ports assists in reducing the induction period interference between adjacent cylinders, but the variation of flow-path causes some unevenness in the mixture distribution between the inner and outer cylinders.

With the four-branch log-type manifold, the sharp branch elbows of the end cylinders tend to bounce the overshoot fuel particles back into the main air stream of the adjacent branch passages. Four-branch streamlined type manifold Fig. The streamlined junctions made between the inner and outer branch pairs as they divide from the central riser are so shaped that equal proportions of mixture are permitted to flow into the adjacent branch passages. The mix- ture is thus transferred from the gallery's central riser to the two end right-angle elbows.

It then into two diverging branch forks which align with the cylinder-head inlet ports. The prin- ciple behind this arrangement is that similar quan- tities of charge exit from the end elbows, and this rge is then equally divided by the streamlined junction so that each inlet port receives the same amount of mixture under all operating conditions.

Note, with this layout both path-lengths for both inner and outer manifold carburettor barrels are primary in action and branches are approximately the same. To compensate for the irregular induction pulse. When one cylinder is being robbed of charge as its inlet valve is closing, the mixture will flow through the balance-pipe from the opp.

The benelit of this arrangement is th the mixture strength and the quality permitted to flow to each pair of cylinders can be individually adjusted to obtain optimum performance. The benefits gained by having twin carburettors is that each one directly supplies the inner and outer adjacent cylinders with the prepared air and fuel mixture, The pronged division of the branches is symmetrical so that equal quantities of mixture enter each inlet port, and the length of the branches before they are joined together reduces the induction Period overlap interference between adjacent in- ner and outer cylinders.

The reduction in induc tion interference between adjacent cylinders is achieved because the mixture entering a branch finds it difficult to reverse its direction of flow. Thus, the beginning of the induction on one cylinder only marginally interrupts the comple- tion of the induction period of an adjacent cylin- der. This arrangement, in effect. Good atomization is, maintained as the mixture flows through the branch passages.

Three-branch manifold with siamese inlet ports with a single carburettor Fig. The flow from cach manifold branch ent jing the cylinder-head will be equally divided at the forked passage junction.

Unfortunately, however, the flow paih to the central pair of cylinders is much shorter than that of the outer pairs of cylinders. In contrast, the flow resistance in the long outer branches will be higher than for the more direct central branch. Six-branch semi-streamlined manifold with a single carburettor Fig. The charge delivery to each cylinder tends to vary, as does the mixture dis- tribution. The variation of mixture strength be- tween cylinders is largely due to overshoot from the intermediate branches causing the accumula tion of fuel droplets in the end branches.

There is also the unequal tract lengths between the riser and the individual cylinder-head inlet ports which, will make cylinder filling uneven at high engine speeds, This is due to the variation in branch length ramming abilities and the difference in flow resistance for the various flow paths.

Six-branch streamlined manifold with a single carburettor ig. The objective of this arrangement is that each branch produces a certain amount of ram charg: ing but due to the variation in branch length the flow resistance between branches varies a8 does the volumetric efficiency between cylinders.

The flow region between the riser and where the branches begin is very important if good mixture distribution is to be achieved at different engis speeds and loads, and under certain operating conditions the spread of mixture strength be- tween cylinders may be large.

Generally, the smooth streamlined flow-path from riser to inlet valve ports offers the least drag to the columns of charge discharging into the eylinders so that high cylinder volumetric efficiency is achieved.

Therefore, fuel is discharged also alter- hately from each carburettor. However, the mix- ture ow path to the outer inlet valve ports is greater than for the intermediate and inner eylin- Ger-head valve ports so that the mixture distribu: tion under certain throttle opening load and specd conditions may give a wider air-fuel ratio variation between eylinders than can be tolerated if lean mixtures are to be burnt. However, the front and rear half: manifolds cannot be made with equal length branch tracts.

If the riser is placed opposite the intermediate cylinder branches the volumetric efficiency in these two eylinders will be slightly higher than for the other four cylinders. Also, a larger amount of the fuel particles will be directed to these intermediate branches compared with the outer and inner branch tracts.

A better position for the riser is between the intermediate and middle branches, as shown in Fig, 5. This is because there is no reversal of mixture flow in the manifolds before the outer eylinder's induction periods commence. On the other hand, the in- fermediate cylinder number? Six-branch split-manifold with uadruple compound carburetors longitudinally positioned throttle spindles Fig.

At some predetermined engine speed and load com- bination, the secondary barrels also open and, since they are situated closer to the cylinders and the ram effect becomes less important.

The mixture strength tolerance between cylinders may have to be relatively large under different throttle open- ings Consequently, equal quantities of charge will enter each cylinder s0 that cylinder volumetric efficiency will be at a maximum.

This arrangement is thus capable of producing the maximum acceleration response time and very high power outputs in the upper speed range of the engine To optimize engine tuning a number of observations should be made. With an in-line sixeylinder engine, there is no overlap between the opening periods of paired cylinders. How- ever, the induction pulses for the end pairs of cylinders occur at regular intervals whereas.

There is no induction period overlap between cylinders with a conven- tional fring order such as Cylinders num- bers 1, 2 and 3 receive mixture from the front twin-birrels and cylinders numbers 4, 5 and 6 are supplied by the rear twin-barrel carburettor Chlinders 1 and 3 are fed individually by the lef.

The overall result with this layout is that the quantity of the mixture entering both the middle and adjacent outer branches of each manifold are relatively similar and uniform throughout the engine's speed range.

However, the flow path distance from riser to the individual cylinder is fabout the same for cach branch pipe and the Separation of each branch reduces the induction period interference between cylinders. A twin-barrel manifold Fig. These manifolds have been designed for use with che popular firing order With the twin riser layout, each feeds three branches either from one cylinder-head Fig.

When branch tract numbers 3 and 4 cross over branch numbers, 2and 5, respectively Fig, 5.

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