Three types of injection system are used with reciprocating internal combustion engines.
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
Jerk pump. A reciprocating fuel pump that meters the fuel and also furnishes the injection
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
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
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
Mf = mass of fuel delivered in one injection
C = average flow coefficient of the nozzle
A = area of the nozzle cross section
ρ= fluid density
p = upstream pressure
go = force mass aceleration constant
t= duration time
θ = duration angle, degrees
N = 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
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
Mf = mass of fuel per injection
A = pump-plunger area
S = effective pump stroke
ρ= fuel density at the pump inlet
ep = 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
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
ep=Φ(sA/aCAn, sbρ/μgo, pogo/ρa2, po/ pi , L/b, A/CAn, R1,..., Rn ). .......................(7-3)
Φ = a function of what follows
A= pump-plunger area
An,= nozzle area
b= plunger bore
C = nozzle-flow coefficient
L = fuel-delivery-line length
s= pump mean plunger speed
a=speed of sound in the fuel
ρ= fuel density
μ= fuel viscosity
po = nozzle opening pressure
pi = pump inlet pressure
A/CAn, L/b = very critical design ratios
R1,..., Rn = remaining design ratios of the system, including pump, line, and nozzle
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
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 valve
2 A/An Leakage decreases as delivery pressure decreases because of
3 po Leakage increases as opening pressure increases
4 L/b Time for pressure waves to traverse line increases as length
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 speed
θ= 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
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-
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
Factor symbol Dimensions
Stream velocity at nozzle exit u Lt-1
Fluid viscosity μ FL-2t
Fluid surface tension γ FL-1
Fluid density ρ ML-3
Air viscosity μa FL-2t
Air density ρa ML-3
Nozzle diameter D L
Design 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
Physical characteristics Symbol Dimensions
Cone angle near the nozzle α 1
Average drop diameter d L
Fraction of drops of a given x 1
size in the spray envelope
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
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
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
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
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.
In 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
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
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
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
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 INJECTION
In 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
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.