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Automobile Division Chairman's Address

Fuel injection in diesel engines P E Glikin, BSc, PhD, CEng, FIMechE Chief Engineer Fuel Injection Equipment, Lucas CAV Limited, Warple Way, London W3 In his address the Automobile Division Chairman develops the theme that fuel injection equipment is the heart of the diesel engine. He explains the task that the fuel injection equipment has t o carry out and how this has been solved in the past. He describes some present-day systems and sets out the problems in optimizing the injection characteristics. Finally he points tofuture trends in thisfield and outlines some recent developments in electronic control of fuel injection. 1 INTRODUCTION

By tradition, the Chairman of the Automobile Division gives his address on a subject on which he has spent most of his working life. In my own case this is diesel fuel injection. The theme that I shall develop is that the fuel injection equipment is the heart of the diesel engine; not only in the obvious sense that if the fuel injection equipment stops, so does the engine, but that most engine characteristics are directly dependent on the fuel injection system. The engine torque curve, fuel cmsumption, smoke, noise and emissions are determined by the quantity and way in which the fuel is injected into the engine combustion chamber. In fact, the development of the diesel engine itself has been linked very closely to what it has been possible to achieve with the fuel injection equipment. In this paper, I should like to describe the task that the fuel injection equipment (FIE) has to carry out; how this task has been solved in the past; the current position; the problems in optimizing the injection characteristics; and finally, likely future trends. 2 THE TASK OF THE FUEL INJECTION EQUIPMENT

In a diesel engine the fuel is not pre-mixed with the air as in the petrol engine, but is injected directly into the combustion chamber near top dead centre, and ignition occurs spontaneously. The FIE therefore has to fulfil three functions : (a) to meter the quantity of fuel to be injected; (b) to arrange that the injection occurs at the right time in the cycle. This implies accurate timing of injection as a function of engine speed and load; (c) to mix the fuel with the air in the short period of time available. This means that the fuel has to be introduced at very high energy levels. The quantity of fuel injected into the engine cylinder determines the torque that is generated and therefore accurate control of the quantity injected is necessary, both from cylinder to cylinder, and over time. An accuracy of timing of the order of L 1" crankshaft is required to optimize the engine performance, smoke, noise and emissions. This address was presented at an Ordinary Meeting held in London on 16 April 1985. The M S was received on 27 February 1985. 78/85 Q IMechE 1985

Finally, to mix the fuel with the air, high injection pressures are needed. The mixing of the fuel and air immediately after injection is one of the primary factors controlling the combustion. This mixing can be achieved by having either a high air swirl, as in indirect injection engines, and a maximum injection pressure of about 300-400 bar, or low swirl, as in direct injection engines, and maximum injection pressures of 450-850 bar. On quiescent direct injection engines, maximum pressures over lo00 bar are required.

3 HISTORY OF FUEL INJECTION EQUIPMENT

In view of the task imposed on the fuel injection equipment, it is not difficult to imagine that this proved one of the major problems for Rudolf Diesel when he constructed his first engine in 1893. His first experiments with injecting oil directly into the engine proved unsatisfactory, and he therefore used compressed air for forcing the fuel into the combustion chamber. This meant using a compressed air cylinder and a compressor for charging it. After considerable developments (1, 2), the single cylinder engine (250 mm bore, 400 mm stroke) was subjected to acceptance tests in 1897, giving the following results: 17.8 Brake horsepower Engine speed, r/min 154 Indicated thermal efficiency, % full load 34.7 half load 38.9 Brake thermal efficiency, % full load 26.2 half load 22.5 Brake specific fuel consumption full load, 0.h.p.h 238

This was the most efficient heat engine built at that time; its eficiency almost doubled that of the average internal combustion engine then available (2), and caused the diesel engine to become a major prime mover, a position which it holds to this day. Figure 1 shows a section of the engine. The integral compressor can be seen, as well as the fuel/air injector at the top of the cylinder head. The air blast injection method was cumbersome and expensive, and efforts were made to replace it by mechanical (or 'solid') injection systems. Nevertheless, air blast

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Early examples of jerk pumps (3)which use principles still employed today are the Fielding pump of 1916 (Fig. 2a) and the Robson pump of 1919 (Fig. 2b). In both of these pumps fuel is sucked into the pumping chamber A through an inlet valve, and then pumped out through a delivery valve. In the Fielding pump fuel delivery starts when recess P cuts off a port in the pump body, and delivery ends when recess E is uncovered by the sleeve C. By moving the sleeve axially, the quantity of fuel pumped can be varied. In the Robson pump, the sleeve has a helical slot machined in it, and this slot uncovers a port in the plunger to spill fuel at the end of injection. To vary fuelling the sleeve is rotated. With this arrangement some means has to be found to prevent the plunger from rotating. The jerk pump principle has become the major method of injecting fuel in all diesel engines. However, in order to give satisfactory results, two conditions had to be satisfied: (a) The pump had to be made to very close tolerances. Since it meters and pumps fuel at high pressure, the clearances between pumping plunger and cylinder had to be of the order of a few microns-this required methods of production that were dificult to achieve; and (b) a means had to be found for dealing with the compressibility of the fuel.

Fig. 1 Section of Rudolf Diesel’s engine, 1897

injection continued to be employed on large engines until the early 1930s. A variety of mechanical systems was pursued, but the one that has found most general acceptance is the ‘jerk pump’ system, in which an injection pump meters the fuel and injects it at high pressure through injectors into the engine combustion chamber.

Although the latter effect may seem obvious to us today, it was an aspect that caused an enormous amount of trouble: engineers found it difficult to understand why, in a simple pumping system with a pump, pipe and nozzle, the flow out of the nozzle did not follow exactly the displacement of the pumping plunger-particularly in relation to the beginning and end of injection. The reason for the importance of the compressibility of the fuel is that high pressures have to be generated in the system, and injection takes place over a relatively short time period. Therefore the time taken to pressurize the system and the wave travel time in the high pressure pipe cause the instantaneous flow of fuel from the nozzle to be substantially different from the rate of displacement of the pumping plunger. This aspect will be treated in more detail later in Section 5.

Fig. 2 Early jerk pumps (a) Fielding pump, 1916 (b) Robson pump, 1919 Proc Instn Mech Engrs Vol 199 No D3

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(a)

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Fig. 3 Early injectors

(a)Thorneycroft, 1908 (b)Ruston, 1909

Two essential features of a jerk pump system are: (a) a nozzle with a preloaded valve to keep it closed until the required pressure has been reached, and to close against cylinder firing pressure, and (b) if there is a high pressure pipe between the pump and the n o d e , a means of ‘unloading’ the pipe at the end of pumping. Reloaded injectors were first used by Thornycroft in 1908 (3,4) as shown in Fig. 3a, and Ruston in 1909 (Fig. 3b). In both cases, a nozzle valve is held on its seat by a preloaded spring, and as the valve lifts under the action of fuel pressure, a larger area is exposed to the pressure, giving the valve a differential action to keep it open. Various means were suggested for unloading the high pressure pipe at the end of injection, but the current method of using a delivery valve at the pump with a

Fig. 5 Detail of pumping element

collar to retract part of the fuel from the pipe goes back to Steinbecker in Germany in 1913 (4). This method was used by the Atlas Diesel Company in Sweden in 1924, and the valve is still sometimes known as the ‘Atlas valve’. Figure 4 shows a design of pump and injector produced by Robert Bosch in 1927 (3). Pumps of this type-having a separate pumping element for each cylinder disposed along a common camshaft-are known as ‘in-line’ pumps. Figure 5 shows the pumping principle in more detail. The pumping sequence is as follows: as the camshaft A rotates it lifts the tappet B and plunger C during the

Fig. 4 Bosch in-line pump and injector, 1927 CQ IMcchE 1985

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pumping stroke until the ports E and H are closed. Continued upward movement of the plunger pressurizes the space under the delivery valve F, until it lifts and allows fuel to be displaced into the high pressure pipe G. The pressure wave that is generated by the action of the pumping plunger, travels along the high pressure pipe at approximately the speed of sound, until a suflicient pressure is built up at the nozzle end to open the nozzle against the action of the spring force. The pumping plunger continues to displace fuel into the high pressure pipe until the helical edge cut into the side of the plunger uncovers the spill port H; the pressure below the delivery valve then begins to drop rapidly, and the delivery valve closes. In order to unload the line, the delivery valve has a collar below the seat which, in effect, retracts a given volume of fuel from the line. The negative pressure wave caused by this retraction travels to the nozzle end and allows rapid closure of the nozzle. The fuel quantity is controlled by rotating the plungers by means of a rod through toothed quadrants that are clamped to the control sleeve. The injector is shown in more detail in Fig. 6. It consists of a nozzle, having an inwardly opening valve,

Fig. 6 Bosch injector, 1927

which is clamped to a nozzle holder that contains the spring. The load on the spring can be adjusted and this load determines the pressure at which the nozzle opens. The peak injection pressure is determined by the nozzle spray hole area and the rate of pumping. What is remarkable about this design of pump is that, with relatively small changes, it is still the construction used on all in-line pumps in production today. Similarly, the design of injector has also only changed in minor respects from that shown in Fig. 6. The adoption of this type of pump and injector, and its production by specialist manufacturers in different parts of the world, was to a large extent responsible for the rapid develop ment of diesel engines for automotive applications from the early 1930s. Although in-line pumps similar to that described above proved very reliable and suitable for the larger multi-cylinder engines used in trucks and industrial applications, they had certain disadvantages on the smaller engines which were being developed for light commercial vehicles and tractors. One drawback was that in order to change the timing of the injection an advance device had to be interposed between the drive from the engine and the fuel injection pump. Since this advance device was subjected to the high pumping torque, it was of necessity fairly bulky and expensive. Also, the force required to move the control rod quickly for speed regulation meant that a relatively large governor had to be used. It was therefore felt that on the smaller automotive applications distributor pumps might be a more cost effective solution. In a distributor pump, a single pumping element serves all the engine cylinders, its outputs being distributed to each cylinder in turn. In the late 1940s and early 1950s a number of fuel injection equipment manufacturers started very active development of such pumps. A pump that was first produced by the Hartford Machine Screw Company (now Stanadyne) in Connecticut, and then, with modifications under licence by CAV in England, is the DPA pump shown in Fig. 7. In this

,Governor link

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Fig. 7 DPA pump Prw lnsln Mech Engrs Vol 199 No D3

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pump, opposed pumping plungers are carried in a distributor rotor, which is driven by the pump drive shaft. As the pump rotates, filling ports in the distributor rotor are first uncovered, and fuel under transfer pressure is forced into the rotor through a metering valve. At part load, the quantity of fuel admitted to the rotor is determined by the position of the metering valve and the transfer pressure. At full load, the maximum amount of fuel that is admitted to the rotor is determined by mechanical stops which limit the plunger movement. With further rotation of the rotor, a delivery port is opened and a ring cam forces the plungers, through rollers and shoes, inwards, displacing the fuel to each of a number of outlets in turn. Unloading of the system is by outward movement of the plungers, controlled by the cam. The fuel delivery at part load is controlled by rotating the metering valve, and this can be done by linking it to a mechanical governor as shown in Fig. 7. Timing can be altered by displacing the cam ring in a circumferential direction by means of an advance piston. This pump first went into production in the UK in 1955, and to date about 17 million pumps of this type have been produced. The success of distributor pumps in the light commercial vehicle and agricultural sectors contributed significantly to the rapid growth of diesel engines in these sectors since the mid-1950s. Another development of the mechanical pumping system which, however, has not had such widespread adoption is the unit injector. It was stated earlier that the compressibility effect of the fuel exerted a major influence on the injection characteristics by changing the shape of the injection diagram and introducing delays due to the wave travel time in the high pressure

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pipe. One way of minimizing both of these effects is to combine the pump and injector and to eliminate the high pressure pipe altogether. A version of such a unit injector made by the DDA Division of General Motors is shown in Fig. 8 (5). The control of fuelling on the unit injector is similar to that on in-line pumps, with the control rods on the various cylinders being linked together. One of the disadvantages of this system is that, apart from possible complications of installation in the cylinder head, a more powerful governor is required than with a multicylinder pump, because of the greater linkage mass. Also, changing the injection timing as a function of speed is rather complicated. This may not be very serious on larger engines, since the natural timing characteristic of a unit injector over the speed range varies less than with a pumppipe-nozzle system. However, where a wide speed range is required, and especially on smaller engines, some speed advance is required for optimum performance. Another form of unit injector is the PT system made by Cummins for their own engines. This differs from the unit injector described above in that the nozzle valve is used as the pumping element. Filling is controlled by metering the amount of fuel admitted to the nozzle valve between injections. Metering is at low pressure through fixed orifices as a function of pressure and time-hence the term 'PT'. 4 FUEL INJECTION EQUIPMENT TODAY

The current situation is that the majority of large trucks use in-line pumps which work on the principle of the pump described in Fig. 4. The size of the pump depends on the duty, and the pumps now have integral governors, boost control devices for turbocharged applications, and advance devices where necessary to change timing as a function of pump speed. Some truck applications use unit injectors, and, at the lighter end of the range, distributor pumps. Since most truck applications have direct injection (DI) engines, maximum injection pressures are generally in the region of 45&850 bar, or, in the case of unit injectors, over lo00 bar. The majority of light commercial vehicles and agricultural applications use distributor pumps. Within the last ten years, there has been a significant increase in the number of diesel engines for passenger cars. This may have been triggered by the first oil crisis of 1973, which placed greater emphasis on fuel economy. The growth in annual worldwide production of diesel engines for passenger cars is shown in Fig. 9. Diesel engines used in cars are indirect injection (IDI) engines, and the great majority use distributor pumps. Maximum injection pressures are of the order of 300-400 bar. Figure 10 shows a comparison, in pie chart form, of the fuel injection pumps used in the various sectors. The size of each circle represents the annual production of diesel engines in that sector in 1982. Although the maximum injection pressures in passenger car applications are relatively low, and the duty on the fuel injection equipment light, in most other respects the demands on the fuel injection equipment are more severe than on truck applications. On passenger cars, the diesel competes with the petrol engine, Proc Instn Mech Engrs Vol 199 No D3

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ing, with a temperature controlled fast idle setting for cold conditions; automatic excess fuel for starting; and key operated electric shut-down. Figure 11 shows a section through a Bosch VE distributor pump (5). On this pump, the single reciprocating and rotating plunger pumps and distributes the fuel to each outlet in turn. The pump has an integral advance device and governor. Figure 12 shows a CAV RotoDiesel DPC pump; this is based on the pumping principles of the DPA pump; but the outward movement of the shoes is limited by leaf springs, and excess fuel can be obtained. The pump also has an integral governor and advance device. The basic designs of nozzles have changed little over the years: multi-hole, fixed hole area, nozzles are still standard for DI engines and variable hole area pintle

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and therefore rapid starting from cold, absence of smoke, low noise and smooth running particularly at idling are essential. Thus, although the basic distributor pumps that are used are simple in principle, considerable additional devices may in practice have to be employed. It is quite common, for example, to have not only speed advance but also light load advance, and cold idle advance; idling and maximum speed govern21

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FUEL INJECTION IN DIESEL ENGINES

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nozzles for ID1 engines. There have been some improvements in the construction of injectors as a whole with the introduction of smaller fitted diameters in the cylinder head, and with the reduction of the moving mass by using low-spring designs (see Fig. 13). More recently for small ID1 engines, mini-pintles and ‘Microjector’ injectors (12) have been developed as alternatives to the standard LC injector (see Fig. 14). The mini-pintle injec-

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Fig. 14 Injectors for ID1 engines (a) standard LC injector (b) mini-pintle injector (c) Microjector injector

tor is a long-stemmed version of the conventional pintle nozzle, but the outside diameter of the injector is reduced. The Microjector, on the other hand, is an outwardly opening poppet injector, and it is possible to construct it with very small overall dimensions-hence its name. It is likely that the small injectors will find greater application on small ID1 engines for passenger cars in the future.

5 OPTIMIZATION OF FUEL INJECTION PERFORMANCE

(a)

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Fig. 13 Injectors for DI engines (a) high spring injector (b) low spring injector (Q IMechE 1985

This section will examine how the rate of injection from the nozzle depends on the compressibility of the fuel, and also the relationship between rate of injection and timing on the one hand and engine performance on the other. Proc lnstn Mech Engrs Vol 199 No D3

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5.1 Tbe foe1 injection system

It was mentioned earlier that one of the major factors affecting the behaviour of the fuel injection system was the compressibility of the fuel. This aspect will now be considered in more detail. By way of example, a fuel system for a typical DI engine is chosen, displacing 60 mm3/stroke at 1000 r/min (i.e. 2000 r/min engine speed) through a hole-type nozzle. Beginning of pumping occurs when the pumping plunger covers inlet and spill ports, and the end of pumping occurs when the spill port is re-opened as in a conventional in-line pump. This is shown diagrammatically in Fig. 15. The pressures, nozzle movement and rate of injection shown in Fig. 15 were calculated using the Lucas CAV computer program known to give good correlation with experimental results. Figure 15

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shows that the pressure at the pump end begins to rise before the spill port is cut off by the pumping plunger and rises relatively slowly. Similarly, when the spill port is re-opened, the pressure takes 5-6" to collapse. These effects are due partly to the restriction of the spill port, and partly to the compressibility of the volume of trapped fuel at the pump end (the larger the volume, the slower the rise and fall of pressure). The high pressure pipe acts as a transmission line, and the delay between the pump and nozzle end pressure diagrams is due to the travel time of the compression wave (at approximately the speed of sound in the fuel) along the pipe. At the nozzle end, there is a further delay until the pressure reaches nozzle opening pressure, when the nozzle needle lifts and injection begins. When the spill wave at the end of pumping reaches the nozzle, the pressure falls and, when the nozzle closing pressure is reached, the nozzle needle drops to its seat and injection ceases. It will therefore be seen that the effect of the compressibility of the fuel is to change the shape of the injection diagram, so that the rate of injection at the nozzle is less than the rate of plunger displacement at the pump, and to delay the injection at the nozzle relative to the events at the pump end in terms of crank angle degrees. Both of these effects are a function of pump speed, and become more pronounced as pump speed is increased. Thus, at maximum pump speed, the ratio of mean pumping rate to mean injection ratesometimes called the 'spread-over ratio'-an be of the order of 2 to 3. At low speeds this ratio is about 1. Ideally, the spread-over ratio should increase only slightly over the speed range, and therefore the application of the FIE is a compromise between high and low speed performance. Figure 15 shows the effect of fuel compressibility in a pumppipe-nozzle system. It is interesting to compare this with the result that would be obtained with a unit injector having the same pumping rate and same nozzle hole size. The total volume of fuel in the unit injector is less than the sum of the volumes at the pump and nozzle in Fig. 15, because an unloading delivery valve is not needed. Figure 16 is the result of the calculation. This shows that there is less delay between start of pumping and start of injection than in Fig. 15, and also that the period of injection is less (i.e. the average rate of injection is higher). It can also be shown that the period of injection changes less over the speed range than with a p u m p p i p n o z z l e system. Therefore, for a given period of injection at maximum speed, it is possible to have smaller nozzle holes; this may have little effect on engine performance at high speed, but would generally give a better low speed performance. Injection rate and timing are two important parameters that have an influence on engine performance, and this will be considered in the next section. In practice, there are other factors that have to be taken into account when applying a fuel injection system to an engine, and these factors may restrict the optimization of rate and timing. For example, the possibility of reopening of the nozzle due to reflected waves at the end of pumping must be avoided as well as cavitation erosion in the system. In addition, hydraulic changes are sometimes made to the system to modify the (Q IMechE 1985

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delivery curve of the pump as a function of speed. The variables that the fuel injection engineer has at his disposal are changes in the pumping rate, delivery valve, rate of spill at end of injection, pipe diameter, nozzle differential ratio and nozzle opening pressure. The final result is generally a compromise between the various requirements, and this compromise becomes more dificult to achieve as the speed range is increased.

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to ignition. The length of the ignition delay is dependent on the instantaneous cylinder compression temperature at the point of injection: the higher the temperature, the shorter the ignition delay. The length of the ignition delay affects the pattern of burning: the fuel injected during the delay period (shown shaded in Fig. 17) mixes with the air during the ignition delay period and forms a pre-mixed flame giving rise to a rapid heat release and rise in cylinder pressure. The fuel injected after the ignition delay burns as a diffusion flame and gives a more gradual rate of heat release. The specific fuel consumption and combustion noise are related to the heat release diagram. The formation of smoke and NO, depend on the local rates of fuel/air mixing as well as the heat release diagram. Specific fuel consumption, smoke, noise and NO, will now be considered in more detail. 5.2.1 Specijic fuel consumption The specific fuel consumption (SFC) is a direct function of the rate of heat release, and would theoretically be a minimum if all the heat were released instantaneously at top dead centre (TDC). In practice, SFC is reduced as the rate of injection (and hence rate of heat release) is increased, but tests have shown (6)that there is a maximum injection rate beyond which no further reduction in SFC is obtained. This rate is known as the maximum useful rate (MUR) and on any particular engine depends on the swirl ratio (higher swirl gives a lower value of MUR). Figure 18 shows the SFC as a function of timing as the injection rate is increased. This shows that as the rate is increased, the timing for minimum SFC becomes progressively later. Proc Instn Mech Engn Vol 199 N o D3

Black smoke arises from the diffusion phase of the combustion. The pre-mixed combustion does not generate black smoke. During the diffusion burning, smoke generation is increased if high temperatures are reached in fuel-rich zones of the spray, and this is most likely to occur shortly after the ignition delay (7, 8). Therefore, better air/fuel mixing (i.e. higher rate of injection or higher swirl) reduces smoke, as does turbocharging, particularly when combined with intercooling. 5.2.3 Noise Combustion noise is principally a function of the rate of rise of cylinder pressure and this in turn is related to the magnitude of the first peak in the heat release diagram (9). Plotting maximum rate of heat release against combustion noise gives Fig. 19.

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Fig. 20 Trade-off curves between NO, and smoke t t . normal rate of injection x - - - x - - - x high rate of injection Q IMechE 1985

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fore not surprising that factors that reduce smoke (e.g. high injection rate) have the opposite effect on NO,. However, it may be possible, by increasing injection rate and retarding timing, to have'better smoke and better NO,, as seen from the trade-off curve in Fig. 20. It should be noted that since NO, is also formed at part load when the smoke is low, it is possible to reduce the total NO, formed during a driving cycle by using exhaust gas recirculation, without affecting the maximum smoke level. Since SFC, smoke, noise and NO, do not change in the same direction when changes are made in rate of injection, timing, etc., optimization of the engine performance is not easy. One way of considering these parameters simultaneously is to plot a set of trade-off curves (10) as in Fig. 21. These show the effect of two rates of injection at different timings on SFC, smoke, noise and NO,. It can be seen that, for this engine condition, the higher rate of injection can, with suitable timing, give the same noise as the lower rate of injection, and better SFC and NO,.

Since the peak in the heat release diagram depends on the fuel injected during the delay period, it follows that the way to reduce combustion noise is by reducing the ignition delay and controlling the rate of injection during the delay period. It should be noted that the total engine noise is the sum of the combustion and the mechanical noise from the engine. On some engines the mechanical noise is of the same magnitude as the combustion noise. On these engines a major reduction of combustion noise is only worthwhile if some means can also be found of reducing the mechanical noise. 5.2.4 Oxides of nitrogen (NO,) Nitrogen oxides arise from the oxidation of atmospheric nitrogen. The amount of NO, formed increase if high temperatures exist in fuel-lean zones of the spray. In this respect, the formation of NO, is the inverse of smoke which. as mentioned before, is formed when high temperatures exist in fuel-rich zones of the spray. It there-

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Fig. 21 Trade-off curves between noise, smoke, NO, and SFC o-o-o normal rate of injection x --- x - - - x high rate of injection (0 IMcchE 1985

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5.2.5 Unburnt hydrocarbons ( H C )

The Origins Of HC have been studied by various researchers*It has been shown (8, 11) that the formation of HC in DI engines arises from three sources: Too-lean-to-burn mixtures These can occur during the ignition delay period. As fuel is injected during the delay period, more and more air mixes with the spray and the mixture strength in the periphery of the spray becomes weaker. The longer the delay period, therefore, the greater the amount of ‘leanlimit’ hydrocarbons. It has been found by experiment that as the delay period is reduced, there is a critical value below which there is no further reduction of HC emission and hence below which the contribution from the lean limit sources becomes negligible. Nozzle sac volume source After the end of the main injection of fuel when the nozzle needle has seated, there is a small quantity of fuel trapped in the nozzle tip. Some of this fuel escapes into the combustion chamber late in the engine cycle and forms HC emissions. It has been found that HC emission from this source correlates with the volume of the nozzle sac plus the volume in the nozzle spray holes. Fuel-rich mixtures late in engine cycle Experiments have shown that when the fuel delivery is increased HC remains relatively constant until a certain value of fuel delivery is reached, after which HC increases very steeply with further increase in fuel delivery. It is assumed that this is due to the formation of fuelrich mixtures late in the engine cycle. For welloptimized engines, this limit is reached above maximum fuelling, and HC emission from this cause should not normally take place.

In summary, therefore, to reduce HC emission, a short ignition delay should be aimed for, and the volume of fuel below the nozzle seat should be minimized. 5.2.6 Particulates

Particulates are made up of a solid phase of soot particles and a liquid phase of unburnt hydrocarbons that condense in the engine exhaust pipe below 52°C. In one investigation (8) on an ID1 and a DI engine, it was found that, over a range of engine loads and speeds, and at different timings, a good correlation was obtained between particulates on the one hand and smoke and HC on the other. The following relationship was found : particulates = 1.0 x smoke + 0.5 x HC g/m3 g/m3 g/m3 This relationship may not hold over a wider variety of engines or fuel types. However, it seems likely that particulates can be reduced by measures that reduce smoke and HC.

5.3 Requirements imposed on FIE Having considered the various engine parameters separately we now require to see if it is possible to derive a list of the ideal FIE requirements. Roo hstn Mech Engrs Vol 199 No D3

It is not possible to present such a requirement in a strictly quantitative way, since differences-between individual engines sometimes require substantially different FIE characteristics, even though the engines may be of similar size and have a similar construction. Nevertheless, certain general ideals for combustion can be specified, and typical characteristics defined. These apply to ID1 as well as DI engines. There are three ideals for optimizing combustion (8). These are: (a) To achieve ignition with a minimum amount of premixed fuel. This is to reduce the lean limit HC, the liquid phase of particulates, and to reduce noise. (b) To mix the fuel after ignition at as high a rate as permissible. This will minimize the burn time and hence reduce SFC. The higher mixing rate in the diffusion phase of combustion will reduce smoke. However, there is a maximum ‘permissible’ rate of fuel/air mixing because as the rate of mixing increases, so do the cylinder temperature and pressure, thus increasing NO, and noise. (c) To optimize the timing of combustion with respect to crank angle across the speed and load range of the engine. These requirements-particularly (cbtogether with various refinements to improve the overall acceptability of diesel engines suggest a need for a more accurate and flexible control of injection than may be possible through conventional hydro-mechanical means alone. 6 FUTURE TRENDS

A field in which there has been very active development over the last few years has been the application of electronic controls to fuel injection systems. Recent developments in microprocessor technology have opened up opportunities for better control of fuel delivery and timing than has hitherto been possible. In particular, it is relatively easy with electronic control to control fuelling and timing as a function of several inputs: engine speed, temperature, boost pressure, etc. Electronic controls of automotive fuel injection equipment have an electronic control unit (ECU) which is generally mounted in the driver’s compartment. Signals of the driver’s demand, speed, etc. are generated by transducers and taken to the ECU, which then controls actuators that vary fuelling and timing. This is shown schematically in Fig. 22. In this section three different fuel injection system controls will be described: the control of an in-line pump, of a distributor pump, and of a unit injector. These examples also show three different design philosophies. In the case of the in-line pump control, a conventional in-line pump is used, and an electronic controller is attached to it. For the rotary pump, the pump was specifically designed for electronic control. In the case of the unit injector, a fundamentally different way of controlling the injection of fuel was adopted. 6.1 Electronic control of pumps for commercial vehicles

A construction of electronic governor for pumps produced by Robert Bosch for commercial vehicles (13) is Q IMechE 1985

FUEL INJECTION IN DIESEL ENGINES

Engine Engine Boost Air inlet speed andl Pedal water temperature pressure temperature TDC position

I1

I

!I

!I

which is mounted in the driver’s compartment. Inputs from various transducers measuring speed, position of control rod, boost pressure, etc. are taken to the ECU. Fuel is supplied to the injection pump and the servovalve by two electrically driven feed-pumps that provide a pressure of 3 bar. Under conditions of starting or emergency shut-down both feed-pumps are in operation; for shut-down, the feed-pumps are reversed and suck fuel out of the pump fuel gallery, thus giving a rapid shut-down. The pressure from the feed circuit can generate a maximum force at the servo-piston of 145 N, giving a mean effective force in either direction of the order of 70 N for moving the control rod. Thus, sufficient force is available for rapid movement of the control rod if the servo-valve is in a position different from the required position. The closed loop control of the control rod is obtained by taking a transducer signal which is proportional to the position of the control rod to the ECU and comparing this position with the required position as a function of speed, driver’s pedal position, etc. (as stored in the memory of the ECU); the difference between the actual and desired position of the control rod produces an error signal which is amplified and operates the solenoid on the servo-valve. It follows from this that the ‘governor’ can control not only the speed, but also the full and part load delivery characteristics of the pump.

1111

Timing control i

Injectors

Fig. 22 Schematic arrangement of electronic fuel injection

control shown in Fig. 23. This shows a conventional in-line pump on the left-hand side, to which the electronic governor has been bolted. The latter consists of an hydraulically-operated piston which pushes the pump control rod to the left against the action of the spring. The fluid pressure on the piston comes from a servo-valve which is operated by a solenoid. The solenoid is controlled by an ECU

:er

5 Speed transducer

Fig. 23 Electronic control of in-line pump (13)

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6.2 Electronic control of pumps for passenger cars and light duty vehicles A distributor pump that has been developed specifically for electronic control by Lucas CAV (14) is shown in Fig. 24. The distributor rotor has opposed radial plungers as in the pumps described in Section 4, but in this case the radial displacement of the plungers is limited by stops throughout the load range. These stops are provided by angled ramps on the plunger shoes, which co-operate with ramps in the drive shaft of the pump. Control of fuelling is obtained by axial movement of the rotor; as the rotor is moved to the left, the available stroke of the pumping plungers during filling is reduced. A transducer measuring rotor position takes a signal to an ECU, and compares the rotor position with the required value which is stored in the memory of the ECU as a function of speed, pedal position, etc. and the error signal is amplified and taken to an electrohydraulic valve mounted on the pump body which controls fluid pressure to the right-hand side of the distributor rotor. Thus, as in the case of the previous example, there is closed loop control of the rotor position and hence the fuelling. A similar control of advance is obtained by the operation of the piston that displaces the ring cam in a similar manner to that described in Section 4. 6.3 Electronic control of unit injectors In the two preceding examples, conventional mechanical pumping principles were used, but with a mechanical control replaced by an electronic one. For the unit injector shown in Fig. 25, both the timing and quantity of each stroke are determined by the operation of an electro-hydraulic valve (15). To make this possible, a Proc Instn Mech Engrs Vol 199 No D3

P E GLIKIN

174

The fact that it is possible to control the injections on a shot-to-shot basis means that very rapid response of fuelling and timing can be achieved. This method of control opens up the possibility of obtaining pilot injections by operating the valve twice in one injection. The full potentialities of this system have not yet been evaluated.

7 CONCLUSIONS

6

5

3 Distributor rotor

5 Cam ring 6 Shoes 7 Outlet

4 Pumping plungers

8 Rotor position transducer

1 Drive shaft 2 aansfer pump

Fig. 24 Electronic control of distributor pump (14)

very fast acting solenoid is required, and a ‘Colenoid’ actuator (16) is used. The pulse to operate the Colenoid actuator is generated in the ECU, and the timing and duration of this pulse control the beginning of injection and the quantity injected. Very consistent shot-to-shot fuel delivery has been obtained with this system (15).

In this paper I have tried to show how the fuel injection equipment has a major influence on the diesel engine characteristics that determine the acceptability of the engine for its particular duty. The basic methods of pumping the fuel and injecting it into the combustion chamber have changed little over the years. The methods of controlling the fuel injection system, however, have undergone considerable change, with greater demands being made on the engine to meet noise, smoke and emission legislation. The application of electronics to fuel injection systems provides much more flexible control of fuelling and timing, makes it easier to take more variables into account, and opens up the possibility of linking the fuel injection system to the vehicle control.

ACKNOWLEDGEMENTS

The author wishes to thank the Directors of Lucas CAV for permission to publish this paper and to acknowledge the help from colleagues at Lucas CAV in its p r e p aration.

REFERENCES 1 Diesel, R. Die Entstehung des Dieselmtors, 1913 (Springer, Berlin). 2 Cummins, C. Lyle. Internalfire, 1976 (Carnot Press, Lake Oswego,

Oregon). 3 Evans, A. F. The history of the oil engine, 1932 (Sampson Low, Marston, London). 4 Burman, P. C. and DeLuca, F. Fuel injection and controls for internal combustion engines, 1962 (The Technical Press, London). 5 Lilly, L. C. R. Diesel engine reference book, 1984 (Butterworth, London). 6 Greeves, G. Response of diesel combustion systems to increase in fuel injection rate. SAE paper 790037, 1979. 7 Khan, I. M., Greeves, G. and Wang, C. H. T. Factors affecting smoke and gaseous emissions from direct injection engines and a method of calculation. SAE paper 730169, 1973. 8 Greeves, G. and Wang, C. H. T. Origins of diesel particulate mass emissions. SAE paper 810260,1981. 9 Russell, M. F. and Cavanagh, C. J. Establishing target for control of diesel combustion noise. SAE paper 770259,1977. 10 Russell, M. F. Recent CAV research into noise, emissions and fuel economy of diesel engines. SAE paper 770257,1977. 11 Creeves, G. et al. Origins of hydrocarbon emissions from diesel engines. SAE paper 770259, 1977. 12 Howes, P.The new CAV Microjector injector. SAE paper 800509, 1980. 13 Zimmermann, K. D. Elektronischer Regler fuer NutdahmugDieselmotoren. XX FISITA Congress, 1984. 14 Glikin, P. E An electronic fuel injection system for diesel engines. SAE paper 850453,1985. 15 Ives, A. P., FraLI, G. and David, P. Electronic fuel injection

Fig. 25 Electronically controlled unit injector Roc lnstn Mech Engrs Vol 199 No D3

equipment for the light duty diesel engine. Int. Congress of Transportation Electronics, 1984. 16 Silly, A. H. Colenoid actuators-further developments in extremely fast acting solenoids. SAE paper 810462, 1981. Q IMechE 1985