Three types of injection system are used with reciprocating internal combustion engines. They are:1.- inlet-port injection . This type of injection is used with spark ignition only, and the sprays can be either continuous or time. In timed injection the process occurs over a limited range of crank angle. With timed inlet-port injection the spray usually occurs while the inlet valve is open.2.- Early cylinder injection. Fuel is sprayed directly into the cylinder during the inlet or compression stroke, but in any case early enough so that evaporation is complete before ignition. This type of injection is used with spark ignition.3.- Later cylinder injection. Fuel is sprayed into the cylinder near top center in such a way that combustion occurs during the mixing and evaporation process. This is the system used both for diesel engines and for spark-ignition engines as a means of achieving charge stratification and detonation control.

FUEL INJECTION FOR DIESE ENGINES

Injection systems for diesel engines are all of type 3. With this type of late injection, the spray must penetrate into air at compression density. This requirement, togheter with the need for a finely divided spray for rapid mixing, requires high injection pressures. In systems of this type, maximum pressures ahead of the injection nozzle range from about 2000 to over 20,000 psi (140 to over 1400 kg/cm2)The technical problems of fuel injection may be divided into the following two categories :

1.- Metering. By metering is meant the supply of the desired quantity of fuel in each injection , at an acceptable rate and over an acceptable crank angle

2.- Spray formation. This term signifies control of the physical characteristics of the spray so as to secure proper mixing of fuel and air, both in time and space.Although these two aspects of fuel injection are not completely independent, it is convenient to treat them separately, as far as possible.

Fuel metering for diesel engines

For purposes of discussion the following definition will be used :Nozzle. The oppening through which the fuel emerges into the air .Nozzle assembly . An asembly of parts including the nozzle, which is removable from the engine as a unit .Needle. A rod, the inner end of which is used to open and close the nozzle passage. The needle is usually spring-loaded. Most, but not all, high-pressure injection system use some form of needle.Opening pressure. The fuel pressure required to lift the needle under test-bench conditions, with the pressure applied gradually.Injection pressure. The pressure on the upstream side of the nozzle at any particular moment. Jerk pump. A reciprocating fuel pump that meters the fuel and also furnishes the injection pressure.Inlet port. The port through which fuel enters the jerk-pump cylinder .

Port closing. The point where the pump port close, that is, the begining of the effective pump stroke .Port opening. The point at which a relief port, a communicating with the pump inlet, starts to open .Effective pump stroke. The length of the pump-plunger stroke between port opening and port closing .Pump duration. The time of crank angle between start and end of the effective pump stroke.Primary pump. A pump that supplies fuel, under relatively low pressure, to the injection pump.Injection quantity. The mass of fuel in one injection.Injection volume. The volume of fuel in one injection, measured at the injection-pump inlet pressure and temperature.Spray duration. The time of crank angle between beginning and end of the spray.Spray penetration. The visible length of the spray, measured from the nozzle , at a given time or crank angle after the start of injection .Spray length. The maximum spray penetration.Cone angle. The maximum conical angle of the spray envelope, with its apex at the nozzle opening. Excluding air injection, which is now obsolete, there are two basic methods of metering in current use ; they are called the common-rail and the jerk-pump systems. The basic principles of these two systems are illustrated diagrammatically in Fig. 7-1. in common-rail system, fuel is supplied at a constant high pressure from a reservoir of large volume compared with one injection volume. The reservoir includes the passages connecting several injection valves, which is the “common rail” from which the system derives its name. If all passages are large compared with the nozzle and if there are no serious dynamic disturbances in the reservoir system, the pressure on the upstream side of the nozzle can be considered constant and equal to the supply pressure at all times. Assuming the pressure difference across the nozzle to be in the critical range, and therefore independent of the downstream pressure, we can write, from Bernoulli´s theorem:

Mf =CAt√2ρpgo = CA √2ρpgo θ/360N

Where:Mf = mass of fuel delivered in one injectionC = average flow coefficient of the nozzle A = area of the nozzle cross sectionρ= fluid densityp = upstream pressurego = force mass aceleration constantt= duration timeθ = duration angle, degreesN = engine revolutions per unit time

It is at once evident that the fuel quantity delivered will decrease with increasing speed unless θ is so increased that θ/N remains constant. Since this requirement introduces mechanical complications, the common-rail system is not often used except on engines designed to run at constant speed. For such engines, Mf is controlled either by varying θ or

by varying the needle lift, which, when small compared with the nozzle diameter, controls the value of C.Another serious disadvantage of the common-rail system is that C is sensitive to wear of the nozzle and of the needle point and to deposits of solid matter on the needle and in the nozzle orifice.Jerk-pump systems. In this system the needle is lifted by the fuel pressure, which is supplied by a reciprocating displacement pump. This pump, in turn, is supplied with fuel at relatively low pressure. The quantity of fuel supplied by this system can be written:Mf = ASρep

Where:Mf = mass of fuel per injection A = pump-plunger area S = effective pump stroke ρ= fuel density at the pump inletep = pump volumetric efficiency defined, as in the case of reciprocating engines, as the volume of fuel delivered, divided by AS.

On comparing Eq. 7-2 with Eq 7-1, a basic advantage of the jerk-pump system becomes evident, namely, that the nozzle characteristics and engine speed do not affect the fuel quantity unless they change the pump volumetric efficiency. Actually, small differences in nozzle characteristics, such as those caused by ordinary wear or moderate deposit formation, have little effect on ep, and thus the jerk-pump system is much less sensitive to such variations than is the common-rail system. Also, with proper design, the effect of speed on ep can be made small over the usual operating range. The net result is that a well-designed jerk-pump system furnishes a nearly constant fuel quantity across the usual operating range at any given setting of the pump-control.On the other hand, the spray duration and the injection pressure are not fixed in jerk-pump systems as they are in the case of common-rail systems; instead, they depend on nozzle as well as on pump characteristics and on operating conditions, as will appear in the subsequent discussion.

Fuel quantity control. The value of Mf in Eq. 7-2 is usually controlled by varying the effective pump stroke S. A common method of accomplishing this result is to vary the angular position of grooves in the plunger in relation to the fuel-inlet port, as illustrated in fig. 7-2, which shows the well-know Bosch system.It is seen from fig 7-2 that, at a given angle of the plunger, the inlet port will be closed when the top of the plunger covers it. From this point, as the plunger rises, the fuel is forced through the check valve into the delivery line. Delivery continues until the helical groove in the plunger uncovers the inlet port. Ideally, at this point the pressure in the cylinder drops to delivery pressure, and all flow through the check valve ceases.Pump volumetric efficiency. Possible causes of volumetric deficiency (1- volumetric efficiency ) in pumps of the type shown in fig. 7-2 include the presence of fuel vapor in the pump cylinder at port closing, back flow through the delivery check valve, fuel and metal elasticity, and leakage.Fuel vapor. Vapor in a jerk pump is usually the result of vapor information in the supply line and can be prevented by avoiding high temperatures in the supply system and supply

line and by using adequate supply pressure. This pressure,usually 25 to 100 psi (2 to 7 kg/cm2), is supplied by a primary pump, usually of the gear variety. Vapor in the jerk-pump cylinder during delivery must be avoided. The use of inlet ports of generous dimensions, together with adequate supply pressure and avoidance of vapor in the pump-supply system are necessary to achieve this result. During the return stroke of the pump there will be some vapor in the pump cylinder until the inlet port opens .Back flow. At the end of fuel delivery, when the port starts to open, the delivery check valve is open. As the pressure in the pump cylinder falls and the check valve closes, it is evident that there may be an appreciable flow of fuel back into the pump cylinder and out through the port. The extent of t his backflow depends on the dynamics of the system, including the mass, lift, and spring constant of the check valve and the phase and amplitude of pressure waves in the fuel line. Obviously, a quick closing of the check valve, without “bouncing”, is desirable. In general, the smaller the mass and lift of the valve, the more satisfactory will be its operation.Compressibility and deflection. With a large volume of fuel and a flexible line, the stroke of the pump might not result in enough pressure to open the needle. This type of action is of course undesirable, but it can be minimized by minimizing the volume ratio and making the mechanical parts very stiff. Volume ratio is here defined as the ratio of fuel volume under delivery pressure to AS, the effective pump displacement volume. In well-designed systems the deflection of the mechanical parts is small enough to be neglected, but the compressibility of the fuel is always an appreciable factor. The importance of a small volume ratio has lead a number of manufacturers to eliminate the injection line entirely and build the pump and nozzle assembly into single units mounted on each cylinder. The great disadvantage of this design is that the fuel pumps for several cylinders cannot be built into a single compact assembly with a single mount and drive connection but must be separately mounted and driven.Leakage. Some degree of leakage in jerk pumps is unavoidable, especially near the beginning and end of the effective stroke, where the leakage distance to the inlet port is short. Leakage can be minimized by using minimum practicable plunger-to-cylinder clearance and maximum feasible length of leakage paths.Mathematics of pump volumetric efficiency. Assuming mechanical deflections to be negligible, an expression for pump volumetric efficiency can be set up in a manner similar to that used for engine volumetric efficiency in volume 1, Chapter 6, which leads to the expressionep=Φ(sA/aCAn, sbρ/μgo, pogo/ρa2, po/ pi , L/b, A/CAn, R1,..., Rn ). .......................(7-3) Φ = a function of what followsA= pump-plunger areaAn,= nozzle area b= plunger boreC = nozzle-flow coefficientL = fuel-delivery-line length s= pump mean plunger speed a=speed of sound in the fuelρ= fuel density μ= fuel viscositypo = nozzle opening pressure

pi = pump inlet pressure A/CAn, L/b = very critical design ratiosR1,..., Rn = remaining design ratios of the system, including pump, line, and nozzle assembly.

The important ratio of fuel volume under delivery pressure to displacement volume of the pump is a function of the design ratios, including especially L/b.In geometrically similar systems the pressure wave pattern will be a function of the mach index, sA/aCAn. The effect of fuel compressibility is included in the mach index and the ratio pogo / ρa2. The latter ratio includes the modulus of elasticity of the fuel from the relation a2 = E/ρ, where E is the bulk modulus of the fuel.Leakage will be a function of pressures and the Reynolds index, sbρ / μgo. When speed or plunger diameter is the only variable, the fraction of fuel that leaks past the piston will decrease as these variables increase because, leakage increases with Reynolds number raised to a power less than one, while fuel flow increases nearly in proportion to s or b2.Figure 7-3 shows typical curves of pump volumetric efficiency against four important variables.The principal factors in each case are probably as shown in the following table:curve Variable in Eq.7-3 Chief influence on ep

1 s Dynamic effects, including backflow through check valve2 A/An Leakage decreases as delivery pressure decreases because of

larger nozzle3 po Leakage increases as opening pressure increases4 L/b Time for pressure waves to traverse line increases as length

increases

Injection rate. Let the rate at which fuel is delivered from the nozzle be dM/dt. With a constant rate during the effective pump stroke(dM/dt)ideal = MN/θ

where M = mass of fuel delivered in one pump stroke N= engine speedθ= injection angle, in radians.

The actual value of the product (dM/dt)(θ/MN) will be a function of the same set of variables as those which control volumetric efficiency (Eq. 7-3). Actual rates of injection versus speed and nozzle area for one particular system are shown in fig. 7-4. in general, the type of curve shown for full-load pump setting at speeds of 750 and 1000 rpm is considered desirable, since the curves are free from needle bouncing and the injection rate is nearly uniform. Secondary discharges, such as those indicated at 1000 rpm, part load, are considered especially undesirable. Certainly freedom from needle bouncing is preferable from the point of view of mechanical durability. Curves giving reasonably uniform rates can usually be obtained over the useful operating range by proper selection of line length and nozzle opening pressure. The most difficult conditions are at low speed, as indicated by fig 7-4a.

It is obvious from fig. 7-4 that the duration angle of the spray may be greater or smaller than that of the pump, and that there is always a considerable lag between the point of port closing and the start of injection. The spray-duration angle can be stretched out indefinitely by reducing the orifice size, as shown in fig. 7-4d. Diesel engines generally have a full-load injection duration in the range 20° to 40° of crank angle (10 to 20 ° of pump angle for 4-stroke engines).

Spray formation

The process of formation of a spray from the discharge through an orifice has been extensively investigated by the NACA (7.10-7.25). Figure 7-5 shows sprays of fuel-oil issuing from a round orifice into air at various densities. To analyze this problem, let us first take the simplified case of steady flow of a liquid fuel through a nozzle under conditions in which evaporation of fuel during spray formation is negligible. In this case it is probable that the factors listed in the following table influence the spray-formation process:Factor symbol Dimensions Stream velocity at nozzle exit u Lt-1

Fluid viscosity μ FL-2tFluid surface tension γ FL-1

Fluid density ρ ML-3

Air viscosity μa FL-2tAir density ρa ML-3

Nozzle diameter D LDesign ratios of nozzle R1................R2 1 The elastic characteristics of the fluid, as expressed by its sonic velocity, probably have little influence on the character of the free spray under steady-flow conditions.The physical characteristics of a spray are usually characterized by one or more of the following measures:

Physical characteristics Symbol DimensionsCone angle near the nozzle α 1Average drop diameter d LFraction of drops of a given size in the spray envelope

x 1

Length of the visible spray y L

By using dimensional analysis, we see that the parameter α can be related to the characteristics of the fluid, air, and nozzle in the following way:α= Φ(uρD/μgo , γ/μu, ρa/ρ, μa /μ , R1,...,Rn)...............(7-5).Similarly, the ratios x, d/D, and y/D will be functions Φ2, Φ3, and Φ4 of the same group of dimensionless ratios. In this case, in addition to the Reynolds number uρD/μgo, we have a number covering the influence of surface tension, γ/μu, and also the ratios of air density and viscosity to those of the fuel.

It is usually more convenient to measure the fluid pressure p upstream from the nozzle orifice than the velocity u. In this case, from Bernoulli´s theorem, √2pgo/ρ can be substituted for the velocity u.Although quantitative experimental determination of these functions is far from complete , some quantitative and many qualitative relationships of great interest have been demonstrated (7.11-7.50). Space limitations permit only the more important relations to be mentioned here.Effect of air density. Figure 7-5 shows the great importance of the density ratio ρa/ρ on the character of the spray. Apparently the interaction of fuel and air is a most important element in spray formation. With air density near zero, the jet of oil does not break up within the length photographed.The effect of ρa/ρ on drop-size distribution is plotted in fig. 7-6. Typical values of this ratio in diesel engines are from 0.015 to 0.05 . Within this range the effect on drop size appears to be small.

Effect of Reynolds number. Figures 7-7 and 7-8 show that increasing the Reynolds number, uρD/μgo, increases the tendency of the fuel jet to break up into small element. The variable used to change the Reynolds number was the oil pressure upstream from the nozzle, which varies the velocity u. In figure 7-7 the air density was only one atmosphere. Figure 7-8 compares spray atomization at different Reynolds numbers at an air density more representative of Diesel-engine conditions.Effect of surface tension. Figure 7-9 shows sprays with different liquids at identical injection pressures and air densities. Since these liquids differ in all their physical properties (table 7-1) it is difficult to assign the observed spray differences to any one term in Eq. 7-5. .However, the alcohol and water sprays have nearly the same density and Reynolds number, so that the difference in their spray patterns are largely attributable to the effect of surface tension on the term γ/μu of Eq. 7-5. Evidently the fact that the surface tension of alcohol is about one third that of water (see table 7-1) is responsible for the far finer atomization of the alcohol spray. In the range of diesel oils used in practice, differences in surface tension are negligible.Effect of nozzle design on spray characteristics. Figure 7-10 shows various nozzle designs investigated by the NACA. The plain-orifice and pintle nozzles shown in this figure give so-called hard sprays, that is, sprays characterized by a dense core of fuel drops surrounded by an outer layer of well separated drops (fig. 7.11a). The other nozzles give soft sprays, such as those illustrated in fig. 7-12b. Soft sprays have no core; they consist wholly of well separated drops.The drop-distribution curves for hard and soft sprays are plotted in fig. 7-12, while fig 7-13 shows drop-distribution curves for plain nozzles (hard sprays) of various diameters. These figures indicate that hard sprays result in finer atomization than soft sprays and that decreasing the nozzle diameter tends to decrease the drop size.

Relation of spray characteristics to diesel-engine performance

In the case of diesel engines this subject has already been dealt with in Chapter 3 (see especially figs. 3-15 and 3-16 and table 3-3), which illustrated the very close relation existing between spray characteristics and performance in one particular combustion chamber.

In general , every diesel combustion chamber is a special case for which the optimum spray characteristics must be worked out by patient testing of various nozzle and injection-system components. The following general relations will usually be found:

1. diesel engines require hard sprays, as from plain-orifice or pintle nozzles. The reason for this is undoubtedly that the soft type of spray does not have adequate penetration into the very dense air contained in the cylinder at the time of injection (14 to 20 tines inlet density). The core apparently breaks up soon after ignition occurs; otherwise combustion of a good part of the spray would come too late in the expansion stroke.

2. in open chambers, in the absence of strong air motion, the spray must be directed to various parts of the combustion chamber by means of multiple orifices in the nozzle or by more than one nozzle in the cylinder.

3. strong air motion, or swirl, is always desirable, because it makes performance less sensitive to spray variations.

4. divided combustion chambers can usually be made to give satisfactory performance with a single nozzle. Inlet induced swirl is not necessary with divided chambers.

5. spray duration at full load should not exceed about 30 crank degrees.6. multiple injections, such as those caused by needle bounce, should be avoided as far

as possible.7. as for the effects of different physical characteristics of the fuel on engine

performance, the fact that substantially the same performance is obtained with gasoline and diesel oil, provided the mass of fuel injected is the same (see fig. 4-21), indicates that when spray characteristics are adjusted for normal diesel fuel, under conditions prevailing in a diesel cylinder during injection, large variations in fuel density, viscosity, and surface tension (see table 7-1) can be tolerated without serious effects on the resultant mixing and combustion process.

Pilot injectionIn an attempt to secure the advantages of a small fuel quantity during the delay period without sacrifice in over-all performance, numerous systems of pilot injection have been proposed. Pilot injection is defined as injection of a small fraction of the fresh fuel early in the injection period, the remainder of the fuel being injected as in conventional injection systems.

It has even been proposed that the period of pilot injection be so adjusted as to correspond to the delay period. The objective of pilot injection is to have only a small fraction of the fuel take part in the period of rapid combustion.Experiments along this line have been made with some success, but the added complication in the injection system, plus the fact that in high-speed engines the whole injection process must be accomplished in an extremely short time, has precluded general application of the principle. In low-speed engines ignition occurs early in the injection process and there is little to be gained by such an arrangement .

Fuel injection for spark-ignition engines

Although a carburetor is usually the simplest and cheapest device available for fuel metering, there may be good reasons for using a separate pumping system for fuel delivery

rather than depending on the pressure difference due to air flow. This method of fuel delivery is called fuel injection. The usual reasons for using such a system include the following : a need for more accurate fuel metering than is practicable with a carburetor ; reducing pressure loss in a venturi ; reducing the need for inlet heating ; improving fuel distribution when separate injectors are used at each cylinder; improving adaptability to the use of electronic control; reducing fuel loss in two-cycle engines (and in four-cycle engines with large valve overlap)by injecting the fuel during the compression stroke, after the valves have closed; and adapting to stratified-charge systems where diesel-type fuel injection is required.Except for stratified-charge engines, the mixture requirements are essentially the same for fuel-injection systems as for carburetors, except that when individual injectors are used at each inlet port less enrichment for acceleration will be necessary.Systems of control may be wholly mechanical, but many include a combination of mechanical, electrical, and electronic elements. In service at the present time are steady-flow systems injecting fuel near the air throttle, steady-flow systems injecting at the inlet ports, and systems with intermittent injection at the inlet ports. High-pressure injection systems with spray ignition are not yet in general use for spark-ignition engines.Most injection systems use air flow as a primary operating parameter, as in a carburetor. Mechanical control systems usually use motion of an air “flap” valve in the intake system as the primary control element. This moves a control valve to regulate the fuel flow, with metal bellows for temperature and pressure signaling. Figure 7-4 illustrates a typical system of this kind.Mechanical systems are usually assisted by electronics where strict control of emissions is required.Figure 7-5 shows an electronic control system using individual injectors at each inlet port. From transducers measuring engine speed, crank position, spark timing, fuel-air ratio (lambda system), jacket temperature, air-inlet temperature, air flow, and throttle position, signals are processed by the microcomputer so as to give the most economical fuel flow rate for each operating regime for all phases of operation, within the limits of satisfactory engine performance and control of emissions. The system illustrated cuts off fuel flow when the throttle is closed and prevents extended periods of idling. In addition to the functions shown, EGR controls and knock limiting can be added.In an all electric-electronic system, air flow may be measured by a hot-wire flow meter rather than a flap valve. Since the hot-wire type responds to air density as well as to air velocity, it furnishes an automatic correction for altitude.Electronic control for diesel engines. Electronic controls are now being used with diesel engines to improve fuel economy and to reduce air pollution, both in vehicles and in stationary applications. The controlled variables may include injection-pump setting, injection timing, and amount of exhaust-gas recirculation (EGR). Inputs to the control system include pump control position and rate of motion; engine speed; injection angle; temperatures of outside air, engine, and fuel; and pressures of atmosphere and in inlet manifold. Smoke emissions are reduced by limiting the rate of pump-control movement and by limiting the fuel-air ratio to the smoke limit for each operating condition. Injection timing and EGR are controlled for best economy consistent with limitations imposed on exhaust pollutants. If practicable filters for particulate emissions are developed, control will be extended to these also.

In the case of stationary diesel engines, less elaborate systems of operational control may be necessary, and there is usually more opportunity for large-capacity exhaust-treatment systems.Electronic control is one of the most important developments in the field of internal-combustion engines since the first publication of these volumes. In motor vehicles, its use is being extended into non-engine areas such as controlling transmissions and braking systems, improving instrumentation, and warning of malfunctions of various elements throughout the vehicle. In stationary and marine power plants, the replacement of mechanical by electronic control systems is resulting in cost savings and improved performance.The reliability and the accuracy of these control systems do no t seem to be limited by electronics, but of course depend on the quality of the transducers that translate the quantities being measured into the appropriate electrical signals and of the devices that translate the outgoing electrical signals into the mechanical or other means required to yield the final result. In spite or their complexity, present systems seem to be quite reliable.

SIZE EFFECTS IN FUEL INJECTIONIn volume 1, chapter 11, it was shown that similar engines used in similar applications theoretically should be operated at the same values of inlet pressure and temperature, fuel-air ratio, and mean piston speed. It was also shown that commercial engines in a given type of service tend to be operated in this manner.Assuming mechanical similitude to include the injection system, and assuming constant values of the engine variables mentioned, with a given fuel and fixed values of po/pi, the injection parameters which will vary with size are fuel-pump volumetric efficiency and spray characteristics.Pump volumetric efficiency. The Reynolds index in Eq. 7-3 sbρ/μgo, will vary as the bore of the pump. This parameter controls relative leakage past the pump piston, which will fall as size increases. A characteristics of many Reynolds number effects is that they tend to be large only in the range of rather low Reynolds numbers. This probably being the case here, one would expect only fuel pumps of very small-bore to show significant changes in relative leakage as size varies.Referring to the problem of metering fuel for spark-ignition engines, the fuel-air ratio as expressed by Eq. 7-7 contains both engine and pump volumetric efficiencies and therefore depends on both engine and injection-system Reynolds indices. As far as metering is concerned, one would expect appreciable Reynolds number effects only in the case of very small pump and cylinder sizes. Within the automotive range of sizes, these effects should be small, in which case Eq. 7-8 applies.Spray characteristics. It has been shown, as in fig. 7-7 and eq. 7-5, that spray characteristics may be strongly influenced by the spray Reynolds number, uρD/μgo. Thus, for any engine with cylinder injection, a change in cylinder size may require changes in design of the spray nozzles, even though the fuel-air ratio required remains the same.