A review of low noise diesel engine design at I.S.V.R.

Journal of Sotmd and Vibration (1973) 28(3), 403-431

A REVIEW OF LOW NOISE DIESEL ENGINE DESIGN AT I.S.V.R. E. C. GROVER AND N. LALOR

Institute of Somtd and Vibration Research, University of Southampton, Sottthampton S09 5NH, England

(Received 3 March i 973)

This paper describes the basic relation between the exciting forces, the vibration characteristics and the emitted noise of a diesel engine. As a result of investigations carried out at I.S.V.R. over the past ten years low noise design principles have been formulated and these have been applied to a variety of engine structures to demonstrate that these principles are feasible in practice. Two such engines, both of which gave overall noise reductions of about 10 dB, are described in detail.

1. INTRODUCTION In recent years the trends of engineering design have been enormously influenced by various legislative and statutory regulations which have been introduced to improve man's standards of safety and comfort. The effects of these regulations are particularly felt in the road transport industry where man is in intimate contact with the engineering product. There are now laws covering every aspect of the construction and operation of the vehicle. One of these deals with the safety aspect and has had marked success in dictating new basic principles for car and commercial vehicle design. Another, covering exhaust emission, has led to great advances in the combustion system design of petrol engines. In the commercial field, in order to meet the legislated power to weight ratio of the vehicle, more powerful engines are required. To increase the power output the engine can be made larger, or can run faster, or its specific load can be increased. Generally, the parameter chosen for increasing the power output of an engine has been the speed, which also, unfortunately, increases the noise [1]. The result is that automotive diesel engine noise has increased by at least 10 dBA over the past 20 years. Such a significant increase in noise has necessitated the introduction of further legislation on road transportation in many countries. The introduction of the high speed diesel engine, however, has offered very great economic advantages since the same power output has been possible from a physically smaller and lighter power unit. The exciting forces that arise from the operation of an engine are difficult to control, but a knowledge of them is essential to the understanding of the engine noise characteristics and hence the noise reduction. The study on which this paper is based was initiated at the Institute of Sound and Vibration Research of the University of Southampton, some ten years ago. It has been a systematic investigation aimed at producing basic design principles acceptable to the motor industry as regards production feasibility for quieter engine structures. The present investigation has covered a large range of both in-line and vee-form automotive engines and a number of 403

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E. C. GROVER AND N. LALOR

experimental low noise engine structures have been built at the I.S.V.R. to demonstrate that this is possible. The authors believe that substantial reductions of noise are possible, but they can only be achieved by radical changes in engine design. 2. MECHANISM OF NOISE GENERATION All noise is the result of vibration. When considering a simple single-degree-of-freedom vibratory system, there are two factors which determine the vibration amplitude: (a) the characteristics of the exciting force; (b) the characteristics of the vibratory s),stem. The emitted noise, however, also depends on the size of the vibratory system in relation to the frequency of vibration. For an engine, which is generally a very complex structure and which is set into vibration by numerous exciting forces of complex nature, the same basic principles apply. It could thus be stated that engine noise is determined by the form, the magnitude and the repetition rate of the exciting forces and the overall structural acoustical response. 3. EXCITING FORCES There are numerous exciting forces present in the diesel engine. They can be classified simply as those due to combustion, and those due to the mechanical operation of the engine. In the latter category, every moving part in the engine produces some force on the structure, but there are some, such as fuel injection equipment and pistons, which produce forces of sufficient magnitude to predominate over the others. 3.1. GAS FORCES Gas forces arise from component frequencies of the cylinder pressure, a.typical time history of which is shown in Figure i(a). Small changes in the shape of the pressure diagram, which have negligible effect on the power output, can produce very marked changes in the noise emitted by the engine. It is clear that the pressure forces resulting from combustion must be expressed by means other than a simple time history so that their noise-producing elements show up clearly. The best way to do this is through a frequency spectrum of a pressure diagram [2]. The frequency spectrum of the cylinder pressure diagram has been drawn in Figure l(b) as a continuous curve through the discrete points that correspond with the terms in the Fourier series. For the 1000 rev/min condition, the discrete points are 8.33 Hz apart for each cylinder of a four-cycle engine. The excitation can therefore be regarded as covering the entire frequency range. Even at a full-speed condition of 3300 rev/min, the discrete excitations are only 27.5 Hz apart. A nearly linear decay rate of the spectrum occurs in the acoustically important region around 1000 Hz. The spectrum divides broadly into three main regions [3]. The first, at low frequency, is a broad peak centred around the fundamental repetition frequency of the gas cycle. Its magnitude is governed mainly by peak cylinder pressure and the width of the pressure diagram. The middle part of the spectrum consists of the linear decay mentioned previously, and in this case has a decay rate of about 30 dB per tenfold increase in frequency. This is typical for a naturally aspirated diesel engine. The slope of this line (which controls the relationship between noise and engine speed) is controlled by the abrupt pressure-rise part of the diesel cycle. This is a function of the ignition delay and the injection rate. A longer

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LO%V NOISE E N G I N E D E S I G N A T I.S.V.R.

ignition delay will produce more severe combustion, which results in a higher instantaneous rate of pressure rise at the beginning ofcombustion. This affects the steepness of the spectrum curve in Figure l(b) and therefore the forcing at the higher harmonics. The final region is composed of a fixed frequency peak due to cylinder pressure oscillations that are set up by the rapid pressure rise at the start of combustion. The oscillations can be clearly seen around the pressure peak of Figure l(a). I

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3.1.1 Effect of combustion system o11cylinder press,re spectrum The combustion system of a diesel engine is responsible for a particular form of cylinder pressure development and depends on many features, such as (i) the actual geometry of the combustion chamber, or its acoustical characteristics, (ii) the inlet system which, to a great extent, is responsible for the induced swirl, (iii) the details of injection (number of sprays, spray angle, hole size) and injection characteristics, and (iv) the operating cycle. The design, research and development work generally attempts to match these closely interconnected factors in an effective and efficient workable system and so it is extremely difficult to single out any one of the factors and discuss its effect on noise.

Comparison of direct attd hldirect hljection combustion systems There are two basic forms of combustion chamber used in diesel engines: direct injection (D.I.) and indirect injection (I.D.I.). The D.I. engine has a single compact combustion chamber accommodated in the piston crown. The mixing of fuel and air, apart from a limited swirl induced by the inlet port and inlet valve design, is provided by some four or more simultaneous sprays from an injector. On smaller engines the physical size of the holes of a multiple orifice nozzle becomes too small to give reliable operation. The precombustion chamber provides a solution where a high rate of swirl can be obtained, giving good mixing of fuel and air, with a single orifice spray. These engines, known as I.D.I. engines, are now made as small as 250 cc/cylinder and can run at speeds as high as 5000 rev/min. Figure 2 compares the basic shapes of the combustion chamber of the D.I. engine and the I.D.I. engine, together with typical pressure diagrams obtained with these two types of combustion chamber.

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The basic form of the pressure diagrams of the two systems is almost identical. The maximum rate of pressure rise can be seen during the initial stages of combustion. The rate of pressure rise decreases with increasing pressure until the peak pressure is reached.

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Figure 2. Relationship between combustion chamber geometry, cylinder pressure and spectra. The exciting propensities of the cylinder pressure developments of the D.I. and the I.D.I. engines, described in the form of cylinder pressure spectra, are also shown in Figure 2. There is one major difference between the I.D.I. and D.I. engines and this is in the nature of the pressure oscillations which are superimposed on the cylinder pressure diagrams. The frequency of these oscillations is different for the two engines. On the I.D.I. engine it is about 2000 Hz, and about 5000 Hz for the D.I. engine, despite the fact that the D.I. engine is of considerably greater size. This phenomenon has been studied previously and is attributed to natural modes of gas vibration in the combustion chamber [4]. The damping in the combustion chamber can be determined from the shape of the peak in the spectrum due to pressure oscillations, and it is found that as the bore size increases the damping becomes less. Comparing damping at both extremes of the size range in terms of

LOW NOISEENGINE DESIGNAT I.S.V.R.

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dynamic magnification factor (Q) we have Q = 30 on the 13 litre/cylinder engine and Q = 5 on the 1.6 litre/cylinder engine. The effects on damping of various combustion chamber configurations were studied on direct injection engines of about l litre/cylinder capacity. It was observed initially that in shallow dished chambers, as shown in Figure 3, the pressure oscillations were far more marked than in the more common toroidal chambers. Q factors of 15 were measured in the dished chamber and 2"5 in the toroidal chamber. This suggested that waves were more highly damped in the toroidal chamber, possibly due to the greater distance for the waves to travel in the clearance between the cylinder head and the base of the chamber. However, subsequent investigation showed that on some engines with toroidal chambers the pressure oscillations can be of very much greater amplitude. The study revealed that engines where the pressure oscillations are marked usually have pistons with recesses for valve clearance which, in effect, increase the space between piston and cylinder head and the damping therefore is considerably smaller. ll.Jll

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Effect of hjection system variables on cylimler pressure spectrmn In any combustion system considerable changes can be made by variations which can be provided by the injection system. This includes not only the geometry of injection (e.g., number of sprays) but also the injected fuel time history (rate of injection) and the injection timing. The injected fuel-time history is purely a design feature of a particular injection system and is determined by the type of pump Used, the cam design and injector design (open hole, pintle, pintle with delay type system, etc.). In all cases the fundamental factor is the ignition delay. Another important factor influencing cylinder pressure development is the injection advance. An example from a direct injection engine is shown in Figure 4. With injection advance the ignition delay increases, causing the initial pressure rise to become greater, and the peak pressure higher. Levels of the cylinder pressure spectra are higher, generally over the whole frequency range. The rate of injection too is significant. A gradual initial injectio n (delay nozzle) can produce pressure diagrams of extreme smoothness in comparison with the more rapid injection characteristics of pintle type nozzles.

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Effect of operathtg cycle on cylinder pressure spectrum There is a fundamental difference between the pressure diagrams of the two engines which considerably affects the exciting propensities, as expressed by Fourier series or spectrum. This difference is illustrated in Figure 5. The two-stroke cycle diagram is wider than that of

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the four-stroke cycle, and this difference has a very profound effect on the level of the harmonic components. When assuming identical pressure diagrams on a degree basis of the two types of engine, and comparing them on the basis of engine repetition rate (i.e., the same power), the following observations can be made: (a) the two-stroke cycle low frequency components are higher for the first five harmonics by a maximum of some 5 dB; (b) from the crossover point at about the fifth harmonic, the levels of the two-stroke cycle spectrum become lower by a maximum of some 6 dB.

3.1.2. Effect of operating parameiers on cylinder pressure spectrum Other factors influencing the cylinder pressure spectrum are the operating parameters: namely, speed and load. 210

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Effect ofspeed. The effect of speed on the cylinder pressure spectrum is shown in Figure 6. By doubling the speed, the shape of the spectrum only slightly changes (except for the high fixed-frequency region), and simply moves to the right by a factor & t w o in frequency. This is because the cylinder pressure diagram keeps, on a crank angle basis, approximately the same shape and magnitude over the engine speed range. Thus, the Fourier series representation will also remain the same except for a corresponding shift in the frequency of each term in the series. This will cause a general increase in cylinder pressure level over most of the frequency range by an amount which depends on the general slope of the spectrum. Effect of load. Figure 7 illustrates that with fixed injection timing the effect of load on cylinder pressure spectra is very small. With increasing load there is some increase of peak cylinder pressure and this affects the low frequency components of the cylinder pressure spectrum--the high frequency components are more or less the same. In some instances, the level of the high frequency components can even be higher at low loads. This is because at low loads the combustion chamber is cooler, the delay period longer and the pressure rise steeper. Conversely, with turbocharging the delay period is reduced and the cylinder pressure development is smoother. This is shown in Figure 9 which also shows the reduction of the high frequency components of the cylinder pressure spectrum. Some fuel injection systems provide for injection retard with reduction of load to control noise at light loads. This characteristic can be incorporated in most types of fuel injection equipment. Figure 8 compares the cylinder pressure diagrams at full load and no load for a D.I. engine fitted with a distributor type injection pump. Also shown are the corresponding cylinder pressure spectra.

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There is a reduction of 10 dB of the high frequency part of the cylinder pressure spectrum at no load and the shape of the spectrum tends to approach that of a petrol engine. 3.2. MECHANICAL FORCES Some forces of mechanical origin are very similar to combustion forces. For example, the time histories of valve push rod force, and injector pump drive torque or injector train push rod force (in the case of an engine with unit injectors), all display sudden increases in magnitude at regular intervals during their cycles of operation. Consequently, the force spectrum is similar in shape to the combustion spectrum although generally lower in magnitude. Of all the impacts that occur with the running engine, piston slap is likely to be the most important due to the high piston mass and relatively large clearance in the cylinder bore [5]. Although the impact forces that occur as a result of this action in a running engine cannot be directly measured, a theoretical estimate of levels can be made from various known parameters. The major slap occurs near t.d.c., where the rate of change of side force on the piston is at a maximum. At this point, the velocity of impact is proportional to the cube root of the engine speed. In order to compare this force with that produced by the gas pressure in the cylinder it must be divided by piston area (gas force = gas pressure • piston area). Estimated impact spectra for the engine from which the gas pressure spectra were obtained are also shown in Figure 6.

3.2.1. Effect of design and operating parameters on the mechanical force spectrum It has been shown that the configuration of the combustion chamber and the operating variables all have a marked influence on combustion and hence on the cylinder pressure spectrum. In a similar way the engine design and operating parameters influence the mechanical force spectrum. However, the actual forces cannot in general be directly measured and hence the effects can only be described in broad terms.

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As previously stated the two main factors influencing the magnitude of piston slap are piston mass and piston to liner clearance. As both of these are functions of cylinder bore the magnitude of piston slap force will increase with engine size. Combustion forces also increase with engine size, due to greater piston area, although in general at a lower rate. It is for this reason that piston slap tends to be the predominant exciting force with very large engines.

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For a given engine the magnitude of the piston slap force will depend on the running clearance. Figure 10 shows a vibration oscillogram taken on the block of a running engine fitted with standard pistons (Figure 10(a)) and thermal expansion controlled strut pistons with substantially reduced running clearance (Figure 10(b)). As can be seen in the second case the block vibration due to piston slap has been almost eliminated. With combustion the effect of speed is to simply shift the force spectrum bodily along the frequency axis as the actual form of the cylinder pressure is not significantly changed. In the case of mechanical impacts, however, the magnitude of the force, which is proportional to velocity of impact, also increases with speed. Therefore, the spectrum will shift vertically as well as horizontally along the frequency axis, as shown in Figure 6.

LOW NOISE ENGINE DESIGN AT I.S.V.R.

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4. VIBRATION RESPONSE CHARACTERISTICS OF ENGINE STRUCTURE All conventionally designed reciprocating engines are basically similar in construction. In-line engines consist of a box divided into compartments. The top of this box is closed by the cylinder head, which is comparatively stiff, and the bottom is closed by a relatively flexible oil sump. Also at the top and about halfway down are horizontal decks that support the cylinders. Although these decks are quite thick, their stiffness is lessened by the cylinder bores. Such a structure is flexible in torsion about an axis parallel to the crankshaft, mainly because the oil sump is not stiff enough to "close the box" effectively. Considered as a beam, the engine is much stiffer in bending in the vertical plane than in the horizontal plane. Vee-form engines consist basically of two "Siamesed" in-line engines although in this case there is little difference between the horizontal and vertical stiffnesses.

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If such structures are excited by a sinusoidally varying force of constant magnitude over a range of frequencies, the vibration amplitude measured at a point will show many resonant peaks. At the frequency corresponding to each peak, the structure will take up a different vibration pattern or mode shape. In general the first mode of vibration occurs at a few hundred Hertz and consists o f a torsional motion about an axis parallel to the crankshaft. At higher frequencies, say, up to 1000 Hz, the engine starts to bend along its length like a homogeneous beam. Above this frequency the structure ceases to behave as a solid body and the panels forming the sides of the bays begin to vibrate independently. At the higher frequencies these modes tend to fall into groups. Each individual mode in a group has the same basic characteristics, but differs in detail from the others. This phenomenon is demonstrated in Figure 11, which shows results of a finite element analysis of a 6-cylinder in-line engine crankcase. In this group there are 12 versions of a particular type of panel mode. The natural frequencies are very close, with only 233 Hz total separation. 4.1. EFFECTOF DESIGN VARIABLESON IN-LINE ENGINE VIBRATIONCHARACTERISTICS Engines are built in a large variety of forms and sizes, and thus it is to be expected that the responses of the different designs to a given cylinder pressure development should show considerable variations.

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E. C. GROVER A N D N. LALOR

Investigations have been carried out at I.S.V.R. on a number of different engines and their characteristics observed in great detail [6].

Wet- and dry-linered enghles When considering the general design of an in-line engine, it can be deduced that the lower part of the crankcase can only have low lateral stiffness. The stiffness is dependent on how the bearing support sections are blended into the cylinder block. Since with the wet-liner design the upper and lower cylinder decks are bored to locate the liner, which is a sliding fit, the stiffness is generally lower. This is shown in Table l, where sump flange bending stiffnesses are compared for two small diesel engine blocks. TABLE 1

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The upper deck is appreciably stiffened when the cylinder head is fitted; the cylinder head is a very stiff box type structure with many internal stiffening members forming the water .passageways, inlet and exhaust porting and bosses for valves, injectors, and fixing bolts.

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The main difference between the two forms ofengine construction is with the low frequency beam bending modes. This is illustrated in Figure 12 where the wet-liner engine shows significantly higher vibration levels over a wide frequency range.

Engines with undershmg mtd skirted crankcases The design of both types of engine is more or less the same, apart from the crankcase of the skirted engine, which is extended below the level of the crankshaft. To illustrate the major differences in the vibration pattern, two 6-1itre, 6-cylinder dry-liner engines are compared. The relevant vibration data are illustrated in Figures 13(a) and (b). The lateral vibration patterns at 250 and 1200 Hz are plotted for the two engines, respectively,

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415

I.S.V.R.

in the upper figures as functions of distance from the top to the bottom of the engine and in the lower figures as functions of distance along the engine at the bottom o f the crankcase on the sump fixing flange. When considering high frequency vibration, it will be seen that over the depth of the crankcase from the lower deck of the cylinder block, vibration increases by some 5 dB on the underslung crankshaft engine. An increase of similar magnitude (4 dB) can be observed down to the crankshaft centreline on the skirted crankcase engine. From this point, however, on the skirted crankcase engine the vibration amplitude increases very rapidly and over the distance of about 4 in a magnification of 12 dB is measured at the bottom flange of the skirt. This indicates that the extended skirt exhibits a "flapping" mode of vibration. Vibtolion occerolion level (dB) 50 60

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Enghte with and without hltermahts crankshaft bearhlg Two 4-cylinder engines are chosen for comparison. There are significant differences of vibration pattern between them which can be attributed to this particular feature of design. One of the main differences was found to be that the fundamental bending frequency ofthe five bearing engine was appreciably higher. On a 4-1itre, 4-cylinder engine with five bearings, this fundamental mode was around 800-1000 Hz. On a three-bearing, 4-cylinder engine the fundamental bending frequency is in the range 400-800 Hz. Another, but far more predominant difference is the panel mode of vibration of the crankcase walls, which is very marked in the three bearing engine, as shown in Figure 14(a) at 1000 and 2000 Hz. This observation has been quoted earlier [7]. The vibration amplitudes are shown for a line along the middle part ofthe crankcase. There are very definite nodal lines along the stiffmembers forming the crankshaft supports at the ends and middle of the engine. Since the unsupported crankcase walls are relatively large, there is a high magnification of vibration level Of the order of 10-15 dB from the edge (stiff section of the engine) to the middle of the unsupported crankcase walls. In the five bearing engine, as shown in Figure 14(b), only a slight magnification of vibration of 2-3.dB is observed on the thin section of the crankcase walls between the stiff members of

416

E. C. GROVER AND N. LALOR

the crankcase in this frequency range (1000-2000 Hz). On the five-bearing engine the thin sections are approximately half the span of those of the three-bearing engine. It would be expected that the thickness of the walls would be the same. Therefore, from simple plate theory, the fundamental frequency of a crankcase wall of the five-bearing engine should be four times greater: that is, around 5000 Hz. As can be seen, this is observed on this engine.

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Enghtes withfour and six cylinders Two engines of the same make, and with identical parts and combustion systems, are chosen for comparison. The patterns of vibration for both the 4- and 6-cylinder engines are shown in Figures 15(a) and (b), in the frequency bands ofthe fundamental bending mode of crankcase vibration. For the 6-cylinder engine the fundamental mode can be recognized in

LOx,V NOISEENGINEDESIGNAT I.S.V.R.

417

the third octave bands of 250, 315, and 400 Hz. For the 4-cylinder engine the fundamental mode is in the higher frequency bands of 630, 800, and 1000 Hz. Ifthe engine is assumed t0 be bending in its fundamental mode as a beam, then the frequency can be determined by simple beam theory. For a given cross section the frequency is inversely proportional to the square of the length. Therefore, if the natural frequency in bending of the 4-cylinder engine is assumed to be 800 Hz, then the frequency of the 6-cylinder engine will be 355 Hz, which is approximately the frequency determined by the experiment. Thus it can be deduced that the lower fundamental frequency of a 6-cylinder engine is due to its greater length. 4.2. VEE-ENG1NEVIBRATIONCHARACTERISTICS The vee engine, which is a compact power unit, has been developed to give much greater horsepower than an in-line engine of the same length. The construction of the vee engine can be considered as consisting of two Siamesed in-line engines, as already stated.

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For this reason most of the vibration modes associated with in-line engines are also present with the vee engine. However, the fact that there are now two cylinder banks introduces another mode which consists of bodily motion of one bank relative to the other [6]. Measurements taken along the cylinder bank at the bank-to-bank frequency indicate that vibration levels are more or less constant. Since the vibration levels measured down the bank increase with distance from the nodal line (corner where cylinder block and crankcase join) bodily movement of the whole cylinder block is indicated as shown in Figure 16. The crankcase also vibrates bodily about this nodal line but in opposite phase. 4.3. VIBRATIONOF ENGINE COVERS The function of the engine covers is to enclose the working mechanisms and to retain oil and water. Normally, they do not add to the structural strength ofthe engine and do not carry any of the working load. Most engines have three main covers: (i) the cover enclosing the valve mechanism, (it) the oil pan, which encloses the bottom of the crankcase and acts as an oil reservoir, and (iii) the cover which encloses the timing mechanism. The basic design requirements of the covers is to provide easy access to the parts or the engine requiring attention during maintenance: that is, the valve gear, timing gear, connecting rod and crankshaft bearings, oil sump, etc. When considering the construction of these covers, there is no reason why they should not be made from any material provided they have

418

E . c . GROVER AND N. LALOR

sufficient strength not to be punctured easily, are able to withstand the working temperature of the engine, and can be made to form an adequate seal with the surface to which they are bolted. Two designsare commonly used: pressed sheet metal (usually sheet metal o f a b o u t 0.036 in thickness) and metal castings. Lightweight materials are usually used, such as aluminium or magnesium) but cast iron is not uncommon. Several factors control the choice of whether the covers are pressed or cast. Steel pressings are cheaper to produce in large numbers, but the cheapness can be offset by the cost of press tools where the cover design is particularly large, or ifit is of a difficult shape to form. Castings are usually used for small production numbers and when some difficulty is experienced in providing an adequate seal with a sheet steel cover which requires a greater number of fixing screws. fO0 90

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Figure 17. Comparative vibration levels of various surfaces of 4-1itre, 4-cylinder diesel engines (engine speed 500 rev/min). Basically the cover is a minor part of the engine and therefore it is usually a cheap item of the engine assembly. For example, a pressed steel valve gear cover costs about 0.1 ~o of the total engine cost. In previous publications by Priede et al. during the past ten years [6, 7], it has been shown that covers generally constitute the predominant source of radiated engine noise and that some reduction of engine noise is always possible by improved cover design. Despite this, there are only a few cases where serious consideration has been given to the design of the cover to reduce noise. In Figure 17 the range of vibration levels ofvarious covers is compared with that of the basic engine structure (cylinder block and crankcase) in third Octave band frequencies for a 4-cylinder, 4-1itrc commercial vehicle engine. On this engine the timing cover and oil pan are aluminium castings and the valve cover is a sheet metal stamping. The maximum vibration levels on all the covers exceed the vibration levels on the basic engine load-carrying structure. For example, on the oil pan, which is attached to the lower part of the crankcase, the vibration levels are higher by up to 8 d B in the frequency range 800-4000 Hz.

419

LOW NOISE ENGINE DESIGN AT I.S.V.R.

It can also be noted that on the two cast aluminium covers the vibration levels are augmented in a higher frequency range (higher natural frequencies of their structure) than with the pressed steel cover where the levels are augmented in the frequency range around 300 Hz. This substantiates [7] that sheet metal covers offer a worthwhile reduction of high frequency vibration. Vibration is transmitted to the covers from the main engine structure casting. The characteristics of the structure vibration will therefore control the characteristics of cover vibration.

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Figure 18 shows the typical variation of high frequency vibration amplitude down the vertical plane for different engine configurations. It can be seen that in each case vibration amplitudes are at a maximum at the bottom of the crankcase: i.e., where the oil sump is attached. As stated above, the sump tends to amplify these already large block vibrations. To a lesser extent the other covers respond in a similar manner, particularly the valve covers in the case of the vee engine. 5. RESULTANT ENGINE NOISE It was stated in section 2 that noise is the result of vibration, and that the way the running engine vibrates will depend on the complex interaction between the exciting forces and the vibration response.

420

E. C. GROVER AND N. LALOR

Although the engine is excited by numerous forces, correlation can be made with a single exciting force alone, for example, the cylinder pressure development, provided its effects are above thoseproduced by all other forces.

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Figure 19. Characteristics of exciting force response and resultant noise of a diesel engine. (a) Cylinder pressure diagram; (b) spectra of existing gas force; (c) overall response of the engine (constant sinusoidal force input of variable frequency); (d) resultant engine noise; (e) increase of noise with speed. The basic mechanism of the combustion induced noise results from the rapid pressure rise (sudden application of load on the piston) which induces transient vibrations in the whole engine structure, and all the outer surfaces subsequently emit the noise. This mechanism can be described in measured terms, as shown in Figure 19. The force is the cylinder pressure development (Figure 19(a)). In this example the rapid rise due to combustion occurs some 10 ~ before t.d.c. Since this force-time diagram is repetitive, its exciting propensities can be fully described by Fourier series (Figure 19(b)), as has already been described. This method of defining the cylinder pressure illustrates graphically the level of exciting force at any particular frequency.

LOW NOISE ENGINE DESIGN AT I.S.V.R.

421

The engine structure response, in terms of noise emitted by the engine structure (engine of automotive size) when subjected to a constant sinusoidal force input of varying frequency, is shown in Figure 19(c). Maximum response is in the frequency range between 1000 to 2000 Hz, which is typical for a machine employing a rigid cast iron structure (the maximum response of engines of larger size is in a somewhat lower frequency range). The resultant noise of the running engine is obtained by the combination (addition of the spectra in dB) of the force and response spectra as shown in Figure 19(d). As this is a linear system [8], changes in the exciting force spectrum will be exactly reproduced in the noise spectrum. As an example the level of the cylinder pressure spectru m can be increased by increasing speed as shown in Figure 19(b). In this case the slope of the spectrum is 30 dB/decade and therefore a 30 dB increase of nois~ with a tenfold increase of engine speed can be expected. This is confirmed in Figure 19(e), where measured overall sound pressure levels (dBA) of the engine are plotted c e r s t t s the logarithm of the engine speed. Any other change to the exciting forces will affect the noise of the engine in a similar way, as illustrated by the noise spectra in Figures 4, 7,8 and 9. 5.1. PREDICTION OF ENGINE NOISE

Thesimilarities in the noise produced by a wide range ofengines indicates that the variation in overall noise (dBA) may bear a fairly simple relationship to the basic engine design parameters. It has been shown that overall engine noise (dBA) increases according to the slope of the cylinder pressure spectrum provided combustion is the predominant exciting force. Tests on I00

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Frequency (Hz) Figure 20. Comparison of noise spectra for various types of engines. (a) Comparison at 1500 rev/min part load of a 4 cylinder engine and a 6 cylinder engine which are otherwise identical and of I lit re/cylinder capacity. x • 4 cylinder 102 dBA; o o , 6 cylinder 103 dBA. (b) Comparison of in-line 6 and V8 engines of same bore at 2800 rev/min full load. o o , In-line 6 cylinder 103"5 dBA; x • V8 cylinder 103-5 dBA. (c) Full load spectra of three in-line 6 cylinder engines of the same bore showing effects of design details. o o , Skirted crankcase aluminium sump 103.5 dBA; A A, skirted crankcase steel sheet sump 103 dBA; x • underslung crankshaft aluminium sump 102.5 dBA. (d) Test bed noise spectra at 2800 rev/min full load for two 1 litre/cylinder V8 engines showing effect of stroke/bore ratio, o o , Stroke/bore ratio 0'8, 106 dBA; x x, stroke/bore ratio 1.06, 103.5 dBA.

422

E. C. GROVER AND N. LALOR

a wide range of engines have also shown that the combined structural and acoustic effects can be related to the cylinder bore diameter. This is illustrated in Figure 20 by the noise spectra o f a numbei- of currently produced engines. As can be seeti only changes in bore size have a significant effect. TABLE 2

Predicted and measured noise levels of automotive diesel engines

Induction

Engine type

V form 8 cyl

Four-stroke N.A.t

In-line 6 cyl

Four-stroke N.A.

In-line 6 cyi

Four-stroke N.A.

V form 8 cyl Four-stroke In-line 6 cyl T.Ch.~ In-line 6 cyl Opposed piston single crank with rocker arms--3 cyl Opposed piston double crankshaft in-line 'Two-stroke 6 cyl Opposed piston double crankshaft in-line 6 cyl In-line 6 cyl

Bore (cm)

Overall noise level (dBA) Rated speed r A (rev/min) Calculated Measured

11-75 11.75 11.43 10"80 13.50 13.97 11.80 11.60 10.48 9.84 9-70 9"84 9.37 7.94 8.89 { 13.97 13.97 12-07 8.26

3300 3300 3000 2800 2600 2100 2600 2800 2800 2800 2800 2800 4000 4000 3500 2600 2100 2200 2400

107-2 107.2 105.5 103.3 107.2 105.5 104-4 104.9 102.7 101.3 101 101.3 104.9 101-3 102 107.3 103.3 101 108.3

108 109 106-5 103.5 109 106 102 105 102.5 103 102 103 107 101.5 101-5 107 102 100 108.5

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LOW NOISEENGINEDESIGNAT I.S.V.R.

423

N.B. For an opposed piston design, the equivalent bore must be used: i.e., B,q = V~B. Table 2 shows a comparison between predicted and measured results for a range o f engines.

6. CONTROL OF ENGINE NOISE The fact that there is very close agreement between measured results and this simple theory demonstrates that all conventionally designed reciprocating engines are basically very similar in construction. This is a universal law because the formula does not distinguish the engine form, structural stiffness, type of covers, or the manufacturer. It can be concluded that all manufacturers make engines to certain standards to withstand the working loads and that radical design changes are necessary if substantial noise reductions are to be obtained. Noise from the structure of a diesel engine is the result of vibration of the outer surfaces. It arises from a system of forces acting from within the engine, which are transmitted through the structure and cause it to vibrate in a series of complex mode shapes. From this analysis, it is clear that there are a number of ways in which noise reductions may be achieved: (i) reduction of the magnitude of the combustion and mechanical forces in the acoustically important frequency range; (it) modification of the transmission paths of these forces by introducing additional damping or stiffness into the engine structure; (iii) reduction of the vibration of outer radiating surfaces of the structure by innovative engine redesign; (iv) reduction of the noise emitted by the outer surfaces by reducing the radiation efficiency, or by means of covers and shields; (v) encapsulation (this is mainly a vehicle installation problem). 6.1. REDUCTIONBY CONTROLOF EXCITINGFORCES Considerable reductions in the noise of an engine can be achieved by the right choice of cylinder bore and operating speed. In general, changes in combustion are made at the expense ofefficiency although turbocharging is a notable exception, in that not only is it advantageous from the noise point of view but also the specific fuel consumption tends to be marginally reduced. As far as the noise resulting from mechanical impacts is concerned reductions are best achieved by keeping the running clearances as low as possible. The use of the autothermic strut piston described in section 3 is an example of this. 6.2.

REDUCTION BY ENGINE DESIGN

Generally, the main structure of an engine constitutes about 40-60 ~o of the external surface area, the remainder:consisting of lightly stressed covers or panels. To effectively reduce the engine noise both these surfaces have to be considered. 6.2.1. Engh~e cotrers Figure 21 shows the reduction of engine noise that could be achieved if the contribution from the covers were eliminated. In the examples shown, engine noise reductions range from 3 to 9 dB in the frequency range that controls the overall noise level. In practice, two basic principles can be applied to low-noise cover designs: application of damping and isolation of vibration. The particular application will dictate the choice. Damph~g. Perhaps the easiest treatment to apply is that ofdamping (provided the cover has nearly flat surfaces). In the simplest case, damping can be applied as an additional material

E. C. GROVERAND N. LALOR

424

to the surface, but this is usually effective only when the panel is thin. It is necessary to fabricate thicker panels from a composite damped material, usually of sandwich construction. Although flat panels of this form are readily manufactured, difficulty arises when deep-drawn stampings are required. In such cases it is sometimes possible to use two or more normal stampings pressed together, relying on interracial friction to provide the necessary damping. 90

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Vibration isolation. Isolated covers are advantageous when the vibration levels at the fixing points are unavoidably high. For example [6], a simple cast valve cover with the main surface area isolated from its fixing flange is particularly suitable for controlling bank-to-bank vibration on a vee-form engine. An isolated rigid fixing flange has the additional advantage of ensuring effective oil sealing. Isolation of covers by means of thick, soft rubber gaskets, washers, and bushed cap screws often present oil control problems. These can be alleviated to some extent by using a composite gasket or washer made of a tensionally stiff thin metal layer bonded between two rubber layers (Figure 22(a)). This is common aircraft practice where oil sealing is critical. The principle ofisolation can be applied to most covers on an engine. The oil pan, however, presents certain difficulties when the rubber is strained by the weight of the pan, or sometimes by the weight of the engine if it is rested on its sump. However, it should be possible to overcome these problems with innovative fail-safe design. The flat shields on the cylinder block can be readily isolated. Where they are loaded by coolant pressure, a restraining lip can be incorporated in the fixing flange, as shown in Figure 22(b). In all cases, for the isolation to be effective, the stiffness of the structure on either side of the bond must be significantly greater than that of the bond material itself. Therefore, isolation of very thin panels is not practical. If the isolated panel is also damped, the noise control effectiveness of the cover is improved. As vibration energy is fed into the covers at their points of attachment, it is essential to ensure that they are at positions of low vibration amplitude. Generally, the best positions are at the stiffer parts of the main structure. Covers should always be fixed directly to a stiffpart

LOW NOISEENGINEDESIGNAT I.S.V.R.

425

of the structure, not via a cantilevered member. Therefore, when designing a structure to take covers, the rigidity of the fixing point should be emphasized. The sketch (Figure 22(c)) illustrates how a typical arrangement can be modified in this way.

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Figure22.(a),(b)Coverisolationtechniques.(c)Panelattachmentdetail. 6.2.2. EnghTestructures There are three basic ways of reducing the noise radiated by the main engine structure: (1) stiffening, (2) damping and (3) reducing the radiating area. Control by stiffenhtg. Radiated noise from the engine structure results from bending of the whole block in the low-frequency range and bending of the individual panels in the highfrequency range. For a constant force input with frequency, the vibration amplitude at resonance is inversely proportional to the natural frequency, and therefore stiffening is effective only to the extent that the natural frequency is increased. Increasing the natural frequency can also be an advantage if it then fails within a range where the exciting forces are less. The forces from combustion tend to decrease with frequency at the rate of 30 dB/decade, as shown by the typical combustion force spectrum in Figure 23. Ideally, ira natural frequency is increased by a factor of two, the level of excitation would be reduced by 9 dB. Although some improvement can be expected by the addition of webs and heavy ribbing, with only a small weight gain, it is difficult to increase significantly the whole range of natural frequencies. This can be achieved only by a general increase of thickness.

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426

E. C. GROVER A N D N. LALOR

A fundamental requirement for the control of noise through increased bending stiffness is a material that makes the increase of natural frequencies possible without increasing weight. The frequency of plates in bending is, to a first approximation, proportional to thickness for all structural materials. By using a low-density metal structure ofthe same weight as a cast iron structure, the greater thickness of the walls would make them much stiffer, and therefore the vibration amplitude for a given force would be much lower. Moreover, the natural frequency would be increased inversely to its density. This is the principle adopted in the experimental magnesium engine [7], in which wall thicknesses have been increased bya factor of 4 to 5. Magnesium was chosen for this experiment. It is considerably more expensive than cast iron when compared weight for weight and has many inherent disadvantages such as liability to galvanic corrosion and low ultimate tensile strength.

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spectra 3 ft from engine. Low-frequency modes of engine vibration depend on the overall bending stiffness of the engine structure. This stiffness is mainly controlled by the cylinder block and head. The crankcase is relatively weak because its general form is that of an open box. It has been found experimentally that an integral bearing assembly (Figure 24(a)) can increase the crankcase bending stiffness with only a small weight increase [6]. This arrangement also has the advantage of restraining the end faces of the engine. The effects on engine noise characteristics are shown in Figure 24(b). Control by damphzg. Control of engine noise by use of structural material with higher inherent damping properties is often considered. Although such materials are available, it is

LOW NOISE ENGINE DESIGN AT I.S.V.R.

427

doubtful whether any real advantage could be gained by their use. Even with a highly resonant structure, such as steel or cast iron, enormous damping is introduced by friction between mating and sliding surfaces. Measurements on a bare engine casting showed the dynamic magnification factor Q of over 200, but the same crankcase in a fully assembled engine showed a Q as low as 16. For a further reduction to be worthwhile, additional damping of the same magnitude must be introduced into the structure. This cannot be achieved practically by attention to the basic load-carrying framework, but significant improvement can be made to individual outer panels. One such possibility is the replacement of these outer panels with a highly damped design. The feasibility of this method was illustrated by the skeleton frame engine [7], where a considerable reduction of high-frequency noise was obtained, by using plates of damped construction screwed to a welded steel framework. Control by reduchlg the radiathlg area. Because it is much easier to control the noise from the covers than from the basic structure, there is a considerable advantage in redesigning the engine so that covers constitute a greater percentage of the total area. It has been shown that crankcase walls are often the main noise-radiating surfaces of the basic structure, and therefore their removal holds a potential advantage. The main functions of the crankcase walls are to help support the crankshaft bearings, retain oil, and provide a location for the oil sump. If the crankcase walls are removed, a bearing beam arrangement can be made to support the bearings, and a deeper oil sump can be assembled to the bottom deck of the cylinder block to retain oil. This larger oil sump will then be attached to a stiff structural member with lower vibration levels than the normal position. 6.3. PRACTICAL LOW NOISE ENGlNE DESIGN

The basic principles of low-noise design have been indicated. By using these in conjunction with the choice of an optimum speed and bore, an entirely new engine could be constructed to take advantage of all these principles. Such an engine must satisfy other constraints such as maximum piston speed, size, weight and of course, the required engine power. An alternative to the complete redesign of an engine is to limit the redesign to the basic structure of an existing engine. This takes advantage of standard production parts such as cylinder head, crankshaft, rods, pistons, liners, camshaft, valve train, and fuel-injection equipment. Two experimental engines, a four cylinder in-line [6] and an 8 cylinder vee-form [10], were built at I.S.V.R. to demonstrate that worthwhile reductions of noise are possible with the use of practical manufacturing techniques and conventional materials. To obtain worthwhile noise reductions all major noise radiating surfaces had to be dealt with and the basic modes of the structural vibration had to be controlled. To achieve this entirely new main engine structures were essential. In-lhle enghle. On this engine, illustrated in Plate l, the crankcase was replaced by a stiff crankframe which supports the crankshaft. The interconnecting beam was introduced to give axial rigidity to the main bearings and to control the fundamental bending mode. The frame was covered by a wrap-around damped sump which was attached to the bottom deck of the block. The design changes on this engine produced a l0 dB reduction in noise over the mid- and high-frequency ranges, compared with that of the engine from which it was derived (Figure 25)." Vee-form engine. On the standard engine on which this design was based, due to an unfortunate choice of de'sign parameters, the frequency of the bank-to-bank mode falls in a range which is both sensitive to the ear and where levels of excitation from combustion and mechanical Origins are high.

428

E.

C.

GRAVER

AND

N.

LALOR

As a major reduction ofengine noise was envisaged it was decided to aim at a design giving reduced amplitude of bank-to-bank vibration by stiffening the roots of the banks. Such a design would be expected to result in a nominal increase in mass accompanied by a small increase in natural frequency. i

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Figure 26. Stiffenedbank details. A major weakness in the standard design is at the camshaft bearings, as shown in Figure 26(a). The section between the banks, which can be considered as a flat plate of 88inch in thickness, is subjected to both compressive and tensile loads and buckling moments when the banks vibrate. Apart from the bearing annuli and the end bulkheads this thin plate is the only tie between the banks at the top of the engine. Moreover its effectiveness is considerably offset by general weakening in this area provided by the tappet guide bores which are closely spaced. On the experimental engine (Figure 26(b)) this plate is of substantial thickness and extends to the full depth of the fuel pump fixing bosses. Although this section is solid for the sake of expediency it could be cored to reduce weight. The banks are further stiffened by a tie below

Plate 1. Photograph of I.S.V.R. crankframe engine with covers removed.

(facing p. 428)

LOWNOISEENGINEDESIGNATI.S.V.R.

429

the camshaft. However, this does not extend the full length of the engine as provision is made to insert the bearing bushes. A tie is also provided between the banks at the front end of the engine. This consists of a rigid cast section joining the banks together at top deck level. This cast tie is cored and utilized as a water transfer duct between the banks. Ideally a similar tie is desirable at the rear of the engine but this would necessitate considerable redesign of the pump drive details. To avoid a bending moment the bosses for the cylinder head fixing studs are extended from the top to the bottom deck of the cylinder block thus ensuring a direct tensile load. This principle is illustrated in Figure 27(a). ///#///////,I/

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430

E.C.

GROVER

AND

N.

LALOR

thickness of the crankcase walls is also not practicable without an enormous weight increase. It was therefore considered that for a low noise engine design the crankcase walls should be removed. The lower decks of the cylinder blocks on the low noise design were extended so that they were cut by the horizontal crankshaft plane thus forming rigid attachment points for the sump (also cylinder block side wall covers) where vibration levels would be low as shown in Figure 28(a). An integral bearing beam was fitted to replace the support previously given to the main bearing caps by the crankcase walls (Figure 28(b)).

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Suitable damped covers were developed to replace those on the standard design. The combined effect of all modifications reduced noise over the entire high-frequency range, as shown in Figure 29. An overall reduction of 9 dBA was obtained.

7. C O N C L U S I O N S

From the investigations outlined in the paper, the following broad conclusions can be drawn. Apart from some operating conditions, mainly at low loads or, if the pressure development is very smooth, the noise of the automotive diesel engine is generally the result ofcombustion. In larger engines mechanical excitation can predominate. The two-stroke cycle offers some advantage over the four-stroke engine if the positive displacement blower noise is controlled. As noise increases with engine speed at a far greater rate than with engine load, turbocharging offers a great potential for reducing engine noise by decreasing maximum engine speed and maintaining the same power output. Pressure oscillations set up in the cylinder can influence the noise of the engine, and in the case of very large engines can be the predominant source of noise. The engine configuration and the details of design for the same stroke-to-bore ratio affect the characteristics of the noise emitted but not its overall level in dBA. More power can be obtained without an increase of noise by the addition of cylinders: four- and six-cylinder in-line and V8 engines produce the same overall level of noise.

LOW NOISE ENGINE DESIGNAT I.S.~/.R.

431

The noise level of the engine is, therefore, already determined in the early stages of design by the choice of the rated speed, the bore, and the combustion system. The predominant noise of an automotive engine is produced by the unstressed engine covers: that is, the valve cover, timing cover and oil pan. For the engines tested, the overall noise was reduced by 3-7 dBA by eliminating cover noise. An effective reduction of noise can be achieved by the use of damped sandwich materials and rubber isolation techniques. F r o m the cost-effectiveness point of view, the use of covers, shields, and isolation is an attractive solution although in the long term redesign of the basic engine structure is potentially more advantageous. However, as design differences in conventional engines have little effect on overall noise levels a radical design change is necessary if major reductions are envisaged. The technical feasibility of low-noise diesel engines has been demonstrated by the use of redesigned structures with suitably modified covers. Without the use of encapsulation techniques, reductions of about 10 dBA have been achieved. ACKNOWLEDGMENTS The authors acknowledge the work of their colleagues, in particular Professor T. Priede, and the many government establishments and engine manufacturers who have sponsored the aspects of research discussed in this paper. REFERENCES 1. T. PRIEDE 1971 Automoti~'e Engine Congress, Detroit, ~tiehigan, January 1971, Paper 710061. Noise in engineering and transportation and its effect on the community. 2. T. PRIEDE1960--1961 Proceedingsof the Institution of i~lechanical Enghteers, London (A D), No. 1. Relation between form of cylinder pressure diagram and noise in diesel engines. Four papers on diesel engine fuel injection, combustion and noise. 3. T. PRIEDE,E. C. GROVERand D. ANDER'rON1968 (March) Diesel Enghteers attd Users Association, Pablication 317. Combustion induced noise in diesel engines. 4. T. PRIEDEand E. C. GROVER 1966-1967 Proceedings of the lnstittttion of Mechanical Enghleers 181 (Pt. 3c), 73-89. Noise of industrial diesel engines. 5. B.J. FIELDINO1968 Ph.D. Thesis, Unicersity of Manchester. Identification of mechanical sources of noise in a diesel engine. 6. T. PRIEDE, E. C. GROVERand N. LALOR 1969 Society ofAutomotice Enghteers, Chicago, Paper 690450. Relation between noise and basic structural vibration of diesel engines. 7. T. PRIEDE, A. E. W. AUSTEtq and E. C. GROVrR 1964-1965 Proceedings of the hsstitution of ~lechanical Enghleers 179, (Pt. 2A) No. 4. Effect of engine structure on noise of diesel engines. 8. D. ANDER'rON and J. BAKER 1973 Society of Automotire Enghteers, Detroit, Paper 730241. Influence of operating cycle on noise of diesel engines. 9. N. LALOR 1971 Symposhtm on Noise Control, Society of Em'iromnental Engbzeers, Norember 1971. Origins of recipro'cating engine noise--its characteristics, prediction and control. 10. S. H. JENKINS,N. LALORand E. C. GROVER1973 Society.ofAutomoth,e Engineers, Detroit, Paper 730246. Design aspects of low-noise diesel engines.