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 t he 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 plunger’s 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 2µ, 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 governor’s 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.
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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. It’s 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.
