Inline Fuel Injection Pumps In 1927 Robert Bosch produced the
first practical diesel pump. This design enabled the newly
developed diesel engine to become a viable engine for many
applications. The method of fuel metering on this initial pump was
port and helix, (highpressure metering). This method of metering
was still being used on most modern injection pumps into l990's.
Bosch has licensed many companies to build these pumps but they all
retain the basic Bosch design principles. Bosch designed pumps are
used on many manufacturers engines. One of the larger pumps in the
Bosch line, the PE/S series has many heavy-duty features, making it
suitable for high-output engines These pumps have been used on
Mack, Navistar and Cummins and countless others throughout the
world. Larger camshafts, plungers, and non-adjustable roller
tappets enable this pump to be used with nozzle opening pressures
of 1,350 Bar, (10,000 to 20,000 psi). The hydro-mechanical versions
of these pumps had many add on features and controls such as a fuel
lift pump, smoke limiter, (aneroid), injection advance unit, and
several different governors. The P size pump is generally used on
engines having more than 200 hp (149 Kwh). To meet increasingly
stringent emissions requirements, manufacturers of injection
equipment are using much higher nozzle opening pressures than
previously. Important information about the pump is stamped on a
plate mounted to the side of the pump. This plate will list among
other items, pump serial number, pump model, and part number.
Component Parts and their Function The pump shown at right is
typical of most inline pumps. The pump housing has a low-pressure
fuel gallery surrounding the pumping elements. This gallery is
sealed from the rest of the pump housing so fuel is available only
to the inlet/spill ports of the pump barrels. An excess supply of
fuel is supplied to the gallery by the transfer pump in most
applications and a return line returns unused fuel to the tank.
This excess flow removes any bubbles that form in the fuel caused
by vibration or aeration and also keeps the pump cool. The
camshaft, (14), is coupled to the engine drive train through
various methods but most commonly a gear train arrangement is used.
The camshaft causes reciprocating movement of the pumping plungers.
The pumping plunger and barrel assembly, (8+4) performs two
functions. It forces fuel past the delivery valve, into the
injection line, and to the nozzle by way of its reciprocating
action, it also controls the quantity of fuel by rotating action.
The roller tappets, (13), ride directly on the camshaft and
transmit its motion to the pumping plungers. The plunger springs,
(11), keep the roller tappets in contact with the camshaft. The
control rack or rod, (15), transmits the action of the governor to
the pumping plunger through the control sleeves, (9).
The delivery valve, (5), seals off the high pressure line from
the barrel during the plungers downward stroke and also reduces
pressure in the line to a predetermined level to prevent secondary
injections in the combustion chamber. Pump Operation The heart of
the inline injection pump is the plunger and barrel assembly, (at
left). This is where fuel at supply pump pressure is pressurized to
injection levels ranging from 140 to 1,350 Bar, (2,000 to 20,000
PSI) and the precise control of fuel delivery is accomplished by
changing the point of register of the helical control edge of the
plunger, (the helix), with the spill/fill port. Because the plunger
fits so precisely in the barrel (approximate clearance is only 2 to
4 microns), there are no sealing rings to retain the injection
pressure as the plunger pumps fuel they seal by the viscosity of
the fuel only. Pumping plungers and barrels are lapped together to
provide this seal. Never interchange a plunger from one barrel to
another. Even the warmth caused by holding a plunger in your hand
can cause it not to fit in its barrel. Plunger and barrels are sold
as a matched set. The pump camshaft lobe provides a constant
mechanical stroke length of the pumping plunger. The plunger is
rotated indirectly by the governor to provide changes in fuel
delivery. The upper edge of the pumping plunger has a vertical
groove which connects the hydraulic pressure above the plunger to
the milled recesses below. Near the top is a helix (or control
edge) this edge provides precise control of fuel delivery by
covering and uncovering the fill/spill ports as the plunger is
driven upwards in the barrel. The barrel may have either one or two
control ports, also called fill/spill ports. Any fuel that does
leak by the plunger is usually collected in an annular groove cut
into the barrel or the plunger and a corresponding duct in the
barrel provides a means of returning this leakage fuel to the
charging gallery. Without this method of returning to the charging
gallery, any fuel that leaks by the plunger would end up in the
engines oil supply and cause it to dilute and lead to engine
damage. The design of these pumps is so precise that fuel leakage
by the plungers is very rarely the cause of diluted engine oil and
if this occurs all other leakage possibilities should be eliminated
before suspecting the pump as the cause. When the pumping plunger
is at its bottom position, fuel from the pump gallery enters
through the fill/spill port/s and floods the area above the plunger
and down the vertical groove to the milled recesses. The plunger is
now forced upward by the camshaft. Initially this upward motion
merely displaces fuel back to the charging gallery because the
fill/spill ports/s is/are still uncovered.
After a short period of upward travel, the plunger leading edge
covers the inlet or fill/spill port/s. This is known as port
closure and is critical to the timing of the injection event.
Continued upward movement will raise the pressure and then force
fuel past the delivery valve into the high pressure line, open the
injection nozzle, and inject fuel into the combustion chamber.
Injection will continue until the plunger has risen far enough to
enable the lower control edge of the helix to uncover the inlet or
fill/spill port/s. At this time, pressurized fuel will rush down
the vertical groove on the plunger and exit through the now open
port/s. This is known as spill and is the end of pressurization;
the pressure will collapse back from the nozzle through the open
port in the barrel and will continue to drop until the delivery
valve closing pressure is reached, typically 2/3 of nozzle opening
pressure, (NOP). The delivery valve will then close sealing the
barrel chamber from the high pressure line. This closing maintains
a residual pressure in the high pressure line so the system is
ready for the next injection to that cylinder. After the end of
fuel delivery, the plunger will continue to be forced upward by the
camshaft, but this movement will not cause any further injection it
merely displaces fuel through the open fill/spill port/s back to
the charging gallery. The plunger stroke can be divided into four
stages. Pre-stroke; this is the movement of the plunger from its
bottom dead centre position to the point of port closure fuel is
merely displaced back to the charging gallery during this portion
of the stroke. Retraction stroke; this is the small portion of the
stroke required to raise the fuel pressure to nozzle opening
pressure or NOP. Effective stroke; this is the plunger stroke while
fuel is actually being delivered to the injector nozzle. Residual
stroke; this is the remaining upward travel of the plunger after
the spill port has been uncovered by the helix until the plunger
reaches its top dead centre position.
The effective stroke of the pumping plunger is the time when
fuel is being sent to the injector. The plungers are milled with a
vertical groove, or it may have cross and centre drillings, and
helical recesses. The function of the vertical groove or cross and
centre drillings is to maintain a constant connection between the
pumping chamber above the plunger and the helical recesses so that
when the helix uncovers the spill port pressure above the plunger
can escape through the drillings or the vertical groove. The length
of plunger effective stroke will depend on where the plunger helix
registers (vertically aligns) with the spill port. Control sleeves
lugged to the plunger permit the plunger to be rotated while
reciprocating. Rotating the plunger in the bore of the barrel will
change the point of register of the spill port with the helix.
Therefore, plunger effective stroke and injected fuel quantity
depends entirely on the rotational position of the plunger. This
rotation is controlled by the requirement of more or less fuel and
has no connection to engine speed or plunger reciprocation. The
plungers rotational position when an engine is in a steady load
condition will not change it will only adjust by operator demand or
load change.
In multiple cylinder engines, the plungers must be synchronised
to move in unison to ensure balanced fuelling at any given engine
load. The control sleeves are tooth meshed or mechanically
connected to a governor control rod or rack, which when moved
linearly, rotates the plungers in unison. This is important. It
means that in any position of the rack, all of the plungers will
have identical points of register with their spill ports, resulting
in identical pump effective strokes. The consequence of not doing
this would be to unbalance the fuelling of the engine that is,
deliver different quantities of fuel to each cylinder causing rough
running or even engine damage. Engine shutdown is achieved by
moving the control rack to the no-fuel position. The plungers are
rotated to a point where the vertical groove will be in register
with the spill port for the entire plunger stroke. The plunger will
merely displace fuel as it travels upward, with no pressurization
possible. In other words as the plunger is driven into the pump
chamber, the fuel in the chamber will be squeezed back down the
vertical groove to exit through the spill port and return to the
charging gallery. Most port helix metering injection pumps use
delivery valves to reduce the amount of work required of each pump
element per cycle. Most delivery valves will have a conical seat, a
retraction piston or collar and flutes to guide it in its bore
while allowing unrestricted fuel flow, without the flutes the
delivery valve could stick open. Delivery valves reduce the amount
of work the pump has to do on the next fuel injection cycle by
isolating the high pressure circuit that extends from the injection
pump chamber to the seat of the nozzle valve and holding it a
pressure somewhat below NOP. Fuel retained in the high pressure
pipes to the injectors between pumping pulses is known as dead
volume fuel. Dead volume fuel is held at a residual pressure below
NOP usually 2/3 of NOP. Delivery valves also help to stop secondary
injections. When the spill port opens in the pumping chamber the
pressure collapses very quickly, the injector nozzle will close
first when its differential pressure is reached, usually 65 to 75%
of NOP. Immediately following nozzle closure the delivery valve
retracts into its holder. As soon as the retraction piston enters
the delivery valve holder the high pressure fuel in the line is cut
off from the open spill port. The delivery valve continues to
retract however until the conical seat contacts the matching cup in
the holder this extra movement allows a minute amount of extra
space for the fuel to occupy thereby lowering its pressure to
residual line pressure. This extra space is known as the swept
volume of the delivery valves retraction piston or collar.
Retraction collar swept volume is matched to the length of the high
pressure pipe to achieve a precise residual line pressure. If the
pressure was retained at close to NOP the rushing fuel slamming
into the closed delivery valve would cause a reflected pressure
wave or surge back toward the nozzle and in certain conditions this
could cause the nozzle to reopen and dribble some fuel into the
combustion chamber which in turn would cause poor fuel economy and
HC emissions. Some delivery valves will have a return flow
restriction valve to further reduce pressure wave reflections or
oscillations in systems where cavitation is an issue.
The delivery valve is held in its closed position on its seat by
a spring and by the residual line pressure. If, for whatever
reason, the residual line pressure value was zero, hydraulic
pressure of around 20 atms, (300 psi), would have to be developed
in the pump element to overcome the mechanical force of the spring.
This mechanical force is compounded when the residual line pressure
is pushing on the delivery valve and establishes the pressure that
must be developed in the pump chamber before it is unseated.
When the delivery valve is first unseated, it is driven upward
in its bore by rising pressure in the pump chamber and it acts as a
plunger being driven upward into the dead volume fuel retained in
the high pressure pipe. By the time the fuel in the chamber and the
pipe unite the pressure will be close to NOP then the injector
nozzle valve (NOP) opens and forces atomised fuel into the engine
cylinder. Pump Housing The pump housing is the frame that encases
all the injection pump components and is a cast aluminium, cast
iron, or forged steel enclosure The pump housing is usually flange
mounted by bolts to the engine cylinder block to be driven by an
accessory drive on the engine gear train. In some offshore
applications of inline, port helix metering injection pumps, the
pump assembly is cradle mounted on its base, in which case, it is
driven by means of an external shaft from the timing gear
train.
Cam Box The cam box is the lower portion of the pump housing
incorporating the lubricating oil sump and main mounting bores for
the pump camshaft. Camshaft main bearings are usually pressure
lubricated by engine oil supplied from the engine crankcase and the
cam-box sump level is determined by the positioning of a return
port. In older injection pumps, the pump oil was isolated from the
main engine lubricant and the oil was subject to periodic checks
and servicing. Camshaft The camshaft is designed with a cam profile
for each engine cylinder and supported by main bearings at the base
of the pump housing. It is driven at 1/2 engine rotational speed in
a four-stroke cycle engine by the pump drive plate, which is
itself, either coupled directly to the pump drive gear or to a
variable timing device. Camshaft actuating profiles are usually
symmetrical, that is, geometrically similar on both sides of the
toe, and mostly inner base circle (IBC the smallest radial
dimension of an eccentric). However asymmetrical (the geometry of
each cam ramp or flank differs) and mostly outer base circle (OBC:
the largest radial dimension of an eccentric) designs are used.
Tappets Tappets are arranged to ride the cam profile and convert
the rotary motion of the camshaft to the reciprocating action
required of the plunger. A retraction spring is integral with the
tappet assembly. This is required to load the tappet and plunger
bases to ride the cam profile and it is necessarily large enough to
overcome the low pressure (vacuum) established in the pump chamber
on the plunger return stroke. This low pressure can be considerable
when plunger effective strokes are long but it does enable a rapid
recharge of the pump chamber with fuel from the charging gallery.
The time dimension within which the pump element must be recharged
decreases proportionately with pump rpm increase.
The Barrel The barrel is the stationary member of the pumping
element; it is located in the pump housing so its upper portion is
exposed to the charging gallery. This upper portion of the barrel
is usually drilled with diametrically opposed ports known as fill
and spill ports that permit through flow of fuel to the barrel
chamber to be charged. Some older systems had only one port this
was changed as pump pressures became higher in order to provide a
hydraulic balance at the spill point to prevent the plunger from
being hammered against on side of the barrel as pressure collapse
occurs. Because it contains the spill ports, both its height and
rotational position in relation to the plunger is critical. Barrels
are often manufactured with upper flanges so that their relative
heights can be adjusted by means of shims and fastener slots permit
radial movement for purposes of calibration and phasing. Plunger
Plungers are the reciprocating (something that reciprocates, moves
backward and forward such as in the action of a piston in an engine
cylinder) members of the pump elements and they are spring loaded
to ride their actuating cam's profile. Plungers are lapped to the
barrel in manufacture, to a clearance close to 2, ensuring
controlled back leakage directed toward a viscous seal consisting
of an annular groove and return duct in the barrel. Each plunger is
milled with a vertical slot, helical recess/es, and an annular
groove. In current truck engine applications, a lower helix design
is generally used but both upper helix and dual helix designs are
sometimes observed. The positioning and shape of the helices
(plural of helix) on a plunger are often described as the plunger
geometry. Plunger geometry describes the physical shape of the
metering recesses machined into the plunger and this defines the
injection timing characteristics. The function of the vertical slot
is to ensure a constant hydraulic connection between the pump
chamber above the plunger and the plunger helical recess/es. A
plunger with a lower helix will have a constant beginning, variable
ending of delivery timing characteristic because the fill/spill
port will always close at the same amount of plunger upward travel
and will open depending on its rotational position.
Upper helix designs will be of the variable beginning, constant
ending type. Double helix designs are designed with both an upper
and a lower helix. Double helix designs will have a variable
beginning and variable ending of delivery; this geometric design
tends not to be often used in highway diesel engines. In the most
common helix designs, plungers have identical helices milled on
both sides of the plunger. These are used in many modern high
pressure injection pumps to provide hydraulic balance to the pump
element at the spill point. This design prevents the side loading
of the plunger into the barrel wall from the high pressure fuel
being suddenly released. A further feature of some plungers is a
start retard notch, or starting groove. Start retard notches are
milled recesses in the leading edge of plungers with lower helix
geometry. The start retard notch is usually on the opposite side of
the vertical slot from the helix and in a position that would
correlate close to a full-fuel effective stroke. The governor of
the injection pump is designed to permit the start retard notch to
register with the spill port only at cranking speeds (under 300
rpm) and usually with the accelerator fully depressed. The
objective of the start retard notch on a lower helix design plunger
is to retard the injection pulse until there is a maximum amount of
heat in the engine cylinder, usually when the piston is close to
TDC. The instant the engine exceeds 300 rpm; it becomes no longer
possible for the start retard notch to register with the spill
port. Rack and Control Sleeves The rack and control sleeves allow
the plungers in a multi-cylinder engine to be rotated in unison to
ensure balanced fuel delivery to each cylinder. Plungers must
therefore be timed either directly or indirectly to the control
rack. The rack is a toothed rod or a notched bar that extends into
the governor or rack actuator housing. The rack teeth or notches
mesh with teeth or levers on plunger control sleeves, which are
either lugged or clamped to the plunger. It must be possible to
rotate the plungers while they reciprocate to permit changes in
fuel requirements while the engine is running. Linear movement of
the rack will rotate the plungers in unison, alter the point of
register of the helices with their respective spill ports, and
thereby control engine fuelling. Comparator bench testing Pump
Calibration Because the plunger and barrel assemblies are matched
lapped sets small differences in delivery volumes occur. Pump
calibration is a test stand procedure in which the plunger helix
point of register with the spill port is incrementally adjusted
either by rotating the barrels slightly or rotating the individual
plungers to alter there position relative to the rack. This ensures
the delivery from each pump element is exactly equal. Pump Phasing
Pump phasing involves setting the port closure dimension of each
pump element so it occurs exactly 120 crankshaft degrees apart,
(for a six cylinder engine). It is performed only on the comparator
bench and can be adjusted by shimming the pump barrels or the
plunger tappets.
Charging Pumps The terms charging pump, transfer pump and supply
pump tend to be used interchangeably, depending on the OEM. The
charging pump is responsible for all fuel movement in the fuel
subsystem. In truck applications using port helix metering
injection, the charging pump is normally a single or double acting
plunger pump, flange mounted to the fuel injection pump and
actuated by a dedicated eccentric on the injection pump
camshaft.
Fuel is pulled under suction from the fuel tank through
hydraulic hose by the transfer pump. A primary fuel filter and or
water separator may also be in series with the pump and tank; or a
more rudimentary pre-cleaner can be integral with the charging
pump. The charging or transfer pump is responsible for producing
charging pressure. It discharges to a secondary filter(s) and then
to the charging gallery in the upper housing of the injection pump.
Charging pressures range from 1 to 5 atms (15-75 psi) depending on
the system. In some cases, a hand primer is fitted to the transfer
pump assembly. Its only function is to prime the system manually
after it has been opened or run dry. Transfer pumps are capable of
delivering far more fuel the engine requires so there is usually a
return line from the charging gallery to return excess fuel to the
tank. This helps to remove any bubbles that form due to aeration
and to keep the fuel cool.
Governor or Rack Actuator Housing Either a governor or rack
actuator housing must be incorporated to a port helix metering
injection pump. This acts as the control mechanism for managing
fuelling. A Diesel engine must use a governor to control the amount
of fuel injected because unlike a gasoline engine there is no
throttle to control the amount of air ingested. Gasoline engines
are managed to run on a stoichiometric fuel ratio of 14.7: 1, but
diesels run with an excess of air at all times. A diesel can have
as much as 1000 times the air required to burn the fuel inside the
cylinder under certain operating conditions. Therefore we must
precisely control the fuel quantity or the engine would quickly
accelerate to self destruction, (1000 RPM per sec). Consider an
engine fuel system that is designed to deliver 185 mm3. of fuel for
each injection pulse at peak torque. While this engine is idling,
(no load), it may need only 18.5 mm3. per pulse, just to keep the
engine running while it is cold (enough to overcome the friction
and inertia of the pistons and crankshaft etc.). As the engine
warms these factors will reduce (less friction etc.), if we supply
the same amount of fuel the engine will run faster and faster until
it disintegrates. A governors job is to sense engine speed and
limit it by cutting the fuel delivery to the amount necessary to
maintain its speed. To run the above engine at 1200 RPM under no
load may require only 20 mm3 of fuel but as load is applied the
requirement will increase perhaps as high as full fuel or 185 mm3.
per cycle. The governor can precisely control fuelling to
accommodate this. The governor will control low idle, (the slowest
speed that the engine will run), high idle, (the maximum engine
speed), and will manage fuelling in between these points based on
driver input and load conditions. Mechanical governors were
originally designed by James Watt in 1788 to control the steam
engine of his day. Mechanical governors use a set of flyweights
that spin in relation to engine RPM. The flyweights always try to
reduce engine fuelling and by that engine speed. Governors match
adjustable spring tension against the centrifugal force generated
by the governor weights. The governor will have a main spring and
an idle spring and in most cases a torque control spring it may
also have a starting spring. The combined effort of these springs
is to push the engine fuel control rack towards full fuel. The main
governor spring tension is affected by the throttle position under
all operating conditions the governor will find a balance between
spring force and weight force to control engine fuelling and
therefore engine speed. Mechanical Governors are set so that at
maximum engine speed the governor weights can overcome the combined
tension of all the spring and hold fuelling to a level that the
engine will not exceed its maximum speed. Mechanical Governors such
as the one above have not been used on highway applications since
the 1990s. Crude attempts were made to control engine emissions on
turbocharged versions of these mechanically controlled inline pump
engines, their prime purpose was to reduce visible smoke emissions.
When a turbocharged engine is accelerated there is always a period
of lag before the exhausted heat energy can spin up the turbo to
increase engine breathing, however on acceleration the rack would
move to full fuel and the available air could not combust the
entire fuel load this would result in a puff of black smoke on
acceleration.
These systems were variably called a puff limiter or smoke
limiter or an aneroid. These devices functioned to delay the fuel
racks travel to full fuel until there was sufficient air to combust
the large fuel load. They consisted of a simple device that
physically limited the racks travel until boost pressure acting on
a diaphragm could overcome spring pressure holding the device
restricting the racks travel. Most of these were on off devices if
boost was below a certain level say 5PSI they held the rack at a
proportion of full travel approximately 60 to 80%. Once boost
pressure exceeded the 5PSI the rack would be allowed full travel.
These aneroids were commonly tampered with by drivers thinking they
could get better fuel economy and performance but remember that any
fuel that exits an engine as black smoke is wasted fuel so the tell
tale signs that an aneroid has been tampered with, that is a puff
of black smoke on acceleration indicates a loss of efficiency
rather than a gain.
A second device was introduced to control rack maximum travel
based on barometric pressure. At higher altitudes the available air
contains less oxygen and therefore cannot oxidize the same amount
of fuel so a barometric capsule limits rack travel in much the same
way as an aneroid however based only on barometric pressure.
. .
Hydro mechanical inline pumps could also be fitted with crude
mechanical timing advance systems that were capable of advancing or
retarding engine timing, (depending on the engine), by 8 to ten
degrees but stricter emission controls spelled the end for these
systems.
The only way that manufacturers could meet the ever stricter
emission control legislation was to devise methods to get greater
control over fuelling and injection timing throughout the operating
range of the engine this was not achievable with mechanical
controls. Inline pump systems were adapted so that they could be
controlled by computer this makes them partial authority managed
engines. The amount of control varied by manufacturer but most
inline pumps were fitted with electronic timing control and
electronic fuel rack position control these changes allowed these
pumps to be used well into the 1990s. One of the most popular
adapted systems was designed by Bosch using PE-7100 and PE-8500
pumps. These pumps featured electronic rack actuators in place of
mechanical governors and timing control devices capable of 20
degrees of timing change also controlled by computer.
In order for the computer to successfully manage these pumps a
variety of sensors were required to relay to the computer details
about engine speed and position, temp, air intake temperature and
boost, throttle position, road speed etc. These signals and more
were input to a computer which then processed the information and
made changes to fuelling amount and injection timing based on
internal fuel and timing algorithms or maps. These maps are
basically a set of pre-programmed instructions in the computers
memory that drive its decision making processes.
The control over fuelling and timing had to be extremely
accurate in order to maintain minimum emissions while not
sacrificing maximum engine performance. The rack actuators that
Bosch used the RE-24 and RE-30 were quite sophisticated they were
equipped with a linear proportional solenoid that was computer
controlled with a pulse width modulated signal that precisely
controlled the current flow to the solenoids magnet. By increasing
the magnetic field the solenoid could overcome return spring
tension and drive the control rack towards a full fuel position.
The stronger the current flow through the solenoids coil the
stronger the magnetic force would become. Its all very fine to be
able to control the racks position by this linear proportional
solenoid however the computer needs verification the desired
position is obtained this was accomplished with a rack position
sensor.
The sensor consisted of a measuring coil as seen above and left
in the low idle fuel position. The coil is energized by the ECM at
5 volts. The coil surrounds a laminate iron core that has a
moveable short circuit ring that travels along the core but does
not contact it. This short circuit ring is attached to the rack so
as the rack is moved by the proportional solenoid the ring moves
along the iron core of the sensor. This varies the strength of the
magnetic field produced by the coil and therefore the induced
signal returned to the ECM. This signal is very precise and is
referenced by the computer control up to 60 times per second so the
exact position of the rack is known at all times. The rack actuator
is by necessity mounted at the rear of the pump which in turn is
attached to the engine and therefore is subject to large amounts of
temperature change. These temperature swings cause changes in
resistance in the position sensor coils winding and could lead to
inaccurate position information.
To combat this problem a reference coil is used that has the
identical sensing coil as the position sensor and a fixed position
short circuit ring. This sends a signal back to the ECM that only
changes with temperature change. This allows the ECM to correct
position data from the position sensor as temperature changes.
The second control item needed to control emissions is timing
with mechanical control of timing very little adjustment could be
made and it was usually up to 8 degrees advance based on speed or 6
to 8 degrees retard based on load depending on engine vocation.
Some systems were slightly more sophisticated but computer control
was needed to ensure compliance. The first thing that was needed
was precise engine speed and position data. Inside the rack
actuator housing a tone or pulse wheel was attached to the back of
the pump camshaft. This is a toothed wheel that turns at camshaft
speed. A speed sensor, (an induction pulse generator), sensor was
installed referencing these teeth and its output frequency would
vary with changing camshaft speed. A second induction pulse
generator sensor called a timing event marker was installed and
this sensor referenced a single notch on the tone wheel marking top
dead centre number 1 cylinder. The second requirement is a physical
way to change timing different methods were used but one popular
method used by MACK was called Econovance. This system allowed
computer controlled changes to engine timing of up to 20 crankshaft
degrees. An initial or static timing set at 4 degrees BTDC could be
limitlessly varied between 4 and 24 degrees BTDC this gave the ECM
great control in terms of managing cylinder pressure and
temperature and therefore emissions.
The Econovance operated as an intermediary device between the
engines pump drive gear and the pump camshaft. It consisted of high
lead screw assembly; this is basically a helically splined sleeve
that was forced along a helical spline that actually drove the pump
camshaft. The sleeve was moved by hydraulic pressure. The ECM
controls a proportional solenoid that in turn controls a hydraulic
spool valve. By precisely controlling the spool through a pulse
width modulated signal the timing could be manipulated by the ECM
to any position within the operating range limits. Eventually even
these advances were not enough to meet the emission standards and
in the mid to late 1990s inline pumps were dropped from the on
highway market. Two main problems associated with these pumps led
to their demise. They could not develop the pressures required
typical pressures developed ranged from 16,000 to 20,000 PSI or
1,100 to 1,400 Bar whereas EUI systems develop up to 30,000 PSI or
2,000 Bar. The second shortcoming stems from the fact that as pump
line nozzle systems the are subject to injection lag and nozzle
closure lag to a much greater extent than an EUI system leading to
fuel droplet sizing and other issues.