WFE. Fircer , —_————— THE INSTITUTE OF MARINE ENGINEERS TRANSACTIONS Vol. 84 PART 9 1972 MARINE MACHINERY FAILURESinstitute on Tuesday, 7th December, 197, Paper read at th MARINE MACHINERY FAILURES B, K. Batten, M.Sc, C.Eng., M.I.Mar.E.* {In 1960, the late Mr. H. N. Pemberton, then chief engineer surveyor to Lloyd's Register of Shipping, introduced a paper to the Institute by say Knowledge of service experience, especially in regard to machinery defect is of considerable value to. both marine engine builders and shipowners, particularly if this ‘knowledge can be utilized in improving design or materials or in preventing mal-operation of machinery”. ‘There is still no better statement with which to begin a similar paper after some ten years more experience, during which the Society has carried out over 1100 specialist investiga tions besides numerous metallurgical examinations of failed components, ‘The present paper covers a range of problems from the unusual to the commonplace and, while by no means a comprehensive catalogue of the Society's world-wide Mr, Batten ities, illustrates there are still many lessons to be learned in design, reliability, and plain engineering sense. IvTRopucTION ‘The technical investigation department of Lloyd's Register of Shipping has the opportunity of seeing numerous interesting ‘machinery failures. Lest this should lead to an attitude of ‘complacency, it is important to realize that each one of these failures is a’ cause of distress to owners, builder or shipper. Invariably, the discovery of the reason for these failures is the result of the combined thought of many people, and often many disciplines, and the necessity of keeping an open mind cannot be ‘over-emphasized. However, one thing is certain, failures do not 4s happen; there i alway a reason: “Acts of God” there may be, but they are invariably clearly God begotten, and the phrase Ys never intended to cover man’s ineptitude oF material i sufficiency. Investigators make mistakes—-the best wll admit it and take ‘a new course of investigation; and without this frank and open approach shared between owners, builders and investigators, pany important facets of machinery breakdown are concealed. (tis a salutary thought that after a further ten years of progress engineers have to contend with largely the same problems—vibra- tion, alignment, material fatigue ete. Until one is prepared to look for failure points in the grass roots of design and fabrication Practic one cannot hope fo se a reduction in these unfortunate Incidents. PROPELLERS, SHAFTING AND STERNGEAR Controllable Pitch Propellers In view of the comparative complexity of controllable pitch propellers, itis natural to expect that the failure rate is higher than for fixed pitch propellers. Considering all. propellers built to Lloyd's Register class between 1960 and 19% with engines developing $000 hp and over, the defects from all ‘causes (excluding contact damage) amounted to 40, an incidence of 11:1 per 100 years of service, Defects have been largely divided among operating mechan- ism failures, blades or their securing bolts and blade seals, “Principal Surveyor of Technical Dept, Lloyd's Register of Shipping: Trans1.Mer 8, 1972, Vol. 4 ‘Two interesting cases investigated concerned failure of a propeller blade pivot and of a control system, Examination of a broken blade pivot on one vessel led to the examination of similar rings on other vessels which were also found cracked (Fig, 1). The rings were made from S.G. cast iron, and the cracks originating from the fillet between the boss and the $6.00 (@) BEFORE MODIFICATION (0) mat MootrcaTiON Fio. 1—Variable pitch propeller blade pivot rings mmMarine Machinery Failures ange had fatigue characteristics, Metallurgical examination showed that while there was extensive galvanic corrosion pitting in the fillets, the fatigue cracks had not originated from these points, but from apparently notch-free surfaces. The structure Of the cast iron contained numerous zones where inoculation hhad not been effective and, while this is normal in fairly large castings, their incidence in'this pivot was higher than ‘normal. Changing these pivots to a carbon steel of 0'32ton/in® UTS. ‘and increasing the fillet radius (Fig. 1b) cured the problem. Failure of the control of a KaMeWa propeller system ‘occurred at the moment a ship was entering locks, A slow ahead ‘movement was given by bridge control, followed, almost im- ‘mediately, by a stop order. It was then noticed that the ship ‘was gathering speed ahead, and excessive ahead propeller pitch ‘was indicated. The officer on the bridge de-clutched the engines and rang for an emergency full astern, This had been anticipated by the watchkeeper who, on seeing the indication of unusually high propeller pitch, changed the pitch manually, too late to prevent bow contact. After clearing the filters, the propeller mechanism was tested by the remote control systems for about 1 hours. Apart from a little dirt in the fine mesh operational element of the ‘change-over filter, the remaining filters in the ine appeared to be clean, ‘There were two systems which could affect propeller pitch, ‘@ pneumatic control system and. an oil hydraulic operating system. To maintain zero pitch, a constant pressure of 14 Ibjin® ‘Was required in the pneumatic system, and should this fall, the propeller would adopt an ahead pitch. In all aspects this system appeared to be effective. ‘The oil supply to the telemotor was bled from the propeller pitch-changing. hydraulic system and passed. through two filtering stages before reaching a control orifice at the entrance to the differential control piston. The balance pressure on this piston was regulated by a needle valve, the position of which ‘was controlled by the pneumatic system, The arrangement was such that a reduction of pressure supply to the differential piston resulted in an ahead pitch movement of the propeller. ‘Analysis of the sludge found in the oil drain tank showed, beside, sand, steel and rust, other particle sizes down from (0-003 in to 6.0001 in which could pass through the 100 micron ‘mesh (0-004 in), and also the pencil filter protecting the first strangler hole. These small soft and friable particles composed of fibres and paint flakes could accumulate at the nozzle entry during manoeuvring periods, and rapidly upset the delicate balance in the hydraulic system. Clearly, the cleanliness of such cil is of paramount importance, as is the necessity to avoid, at design stage, areas where build-up of contaminants and restriction of oil flow could occur. MAIN SHAFTING AND STERN GEAR While the purist may regard shafting only m relation to propulsion, it is an integral part of the machinery of a ship, and failures occurring in this area can clearly not be ignored, Whirling of shafting occurs in certain shafting systems, especially if weardown or misalignment has reduced or removed the load from the forward stern tube bearing. In one such vessel where whirling had been measured to occur within the speed range of 70 to 90 revimin, the screwshaft was found eracked both in the keyway and under cavitation damage on the liner. This liner had been sand cast with a number of core plugs fitted, Fig. 2, one of which had presumably come out, as the surface of the tailsaft was exposed through the eroded bottom of the core hole, ‘A paper on this subject was presented to the Institute in Novembert, Excessive weardown of port and starboard after sterntube bearings on a twin screw passenger ship led to the Society checking the alignment conditions of both sets of shafting using strain gauges, and measuring the dynamic bending stresses in ‘one of the taiishafts whilst in service. ‘The change in amplitude and phase of the first. order bending which was coincident with certain pitching motions of the ship indicated a change in the vertical position of the screw- shaft after point of support. The arbitrary selection of the after mm. Fro. 2—Pitted tailshaft liner point of support is significant, as it affects all subsequent align- ‘ment_and frequency calculations. It is the Society's normal practice, based on experience, (0 assume this point to be one third along the after sternbush measured from the after end. The measurements on this ship suggested that there was inter ‘ittent excessive loading of the after end of the after sterntube bearings due to deflexion of the bossing arms in certain operating and sea conditions, which could have been the cause of excessive ‘weardown, and could, under certain running conditions, lead to short term whirling of the screwshaft. The calculated bodily displacement of the shaft to produce the bending stress measured earlier in the screwshafts was approximately 0-165 in and it was ‘suggested the cantilever movement of the bossing arms might be of the same order. Vibration stresses had been measured in bossing arms at sea ‘on previous occasions, but no cleat relationship between deflexion and stress had been’ established. It is reasonable to. assume, however, that shaft deflexion is not caused by a parallel vertical ‘movement of the bossing structure, but by a hinged cantilever action with maximum deflexion atthe after end. This would mean the longitudinal axis of the bearing being angularly displaced relative to the shaft axis resulting in a periodically reduced effective load-carrying bearing area. ‘The strain gauge investigation was divided in two phases: 1) Jacking the port bossing arm to determine the relation between bossing deflexion and bending stresses in the tailshaft and bossing arm; 2) Measuring the stresses induced in the bossing structure under a ‘given deflexion, and interpolation. of the ‘measured stress, level under operating conditions to derive the associated shaft vibratory bending deflexion, The results showed first that the bossing defiexion, both vertical and horizontal, increased linearly with load, suggesting the bossings were behaving as simple elastic beams. Trane MarB., 1972, Vol $4Marine Machinery Failures ‘The bending stresses in the tailshaft did not increase linearly with deflexion; in fact, there was a reduction in stress in the horizontal plane due to an inboard movement of the shaft contrary to the outboard movement of the bossing arm. This suggested that as the after end of the sterntube was lifted, the relaxation of weight on the forward sterntube bearing allowed the shaft t0 move across (inboard) within the clearance of the beating, and with the reducing rate of increase of vertical strain, that the shafl, which had been turning ahead (outboard) had, in fact, climbed the side of the dry bearing. ‘Calculations were now made of the theoretical increase in shaft bending moment with bossing defiexion, and while the experimental results showed the shaft moving horizontally after a vertical bossing deflexion of about 1 mm, the calculations showed that for the same deflexion the forward sterntube bearing would be, in fact, unloaded and the shaft free. ‘Comparison of these measurements with readings of shaft ‘bending moment variation taken at sea, showed that total bossing deflexions over 5 mm were being experienced, correspond- ing to a maximum stress of = 8 tons/in* on the inside of the bossing arm. ‘Cracked liners due to material fault or, more rarely, arduous operating conditions, sill acount for &! number of talsaft ‘Another large passenger liner completed in April 1969, reported after only six months’ service, excessive weardown on the port aft sternbush. The rate of wear was over ten times as ‘much as on the starboard sternbush, being 0-140 in on the port side as compared with 0-012 in on the starboard side in the same period. Jacking the shafts to measure the reactions on the after plummer bearings, showed, as expected, a significant increase in load on the port aft plummer beating, and a calculated ‘eduction in forward sternbush load. By February, 1970, the port after plummer bearing load reached 700000 Ib, whereas that on the starboard after plummer bearing remained at 41 500 1b against originals of 44 500 Ib and 43900 Ib respectively. Before drydocking, measurements of tailshaft bending stresses were taken at sea which showed a predominantly first ‘order stress vibration with smaller amplitude changes associated with propeller blade forces. Bossing vibrations. were slight ‘A high frequency stress variation in the port shaft was felt to ‘be due to whirling of the muff coupling between tailshaft and the first intermediate length, ‘The port tailshaft was drawn. The liner was found fractured in many places in an entirely random pattern, (Fig. 3), and the synthetic resin bush severely overheated and charred thus blocking all water channels. A metallurgical examination showed that the liner had failed under cyclic thermal stressing and that Fio. 3—Thermal cracking of a liner Trans.1.Mor., 1972, Vol 8 the tailshaft itself was fractured under the liner. ‘The most likely causes of the bush failure were: a) Water starvation due to inadvertent shutting of the service stop valve, though both shafts were supplied from a common line, branched through unmistakably clear gauge glasses and any serious water shortage would, certainly have damaged the forward end shaft seal; by Hygroscopic swelling of the bush material producing. insufficient bearing clearance, though both shafts were stated to have had the same initial clearance. It is certain that the blockage of the stern bush water ‘channels must have occurred early in the ship's life. The continual stopping and starting characteristic of a cruise ship could have induced severe thermal cycling requisite to cause cracking of the liner. If the correct cause is obscure, the moral is obvious; ‘gauge glasses and thermometers etc., however remote from the hhub of activity in the comfort ofthe control room, are hot there just for amusement; they are meant to be used intelligently. In 1966 the Society’s gearing rules were modified to take account of the greater use of surface-hardened gears and to bring oil-engine driven gears within the rule framework. ‘Since that time numbers of gearing failures have been investigated by the Society. Two tankers propelled by twin Diesels developing a total of 100080 bhp, suffered main wheel rim fractures, and whereas a material fault was considered at the time to be the cause of failure, it was decided to oblain a clearer picture of tooth load distribution. Short-base electrical resistance {ype strain gauges at the forward and after end of the face width in the main wheel tooth root radii showed bending strains as the teeth passed through the port and starboard meshes in succes- sion, Gauges were also placed on the rim to detect meshing strains. Signals from all these gauges within the gear case were brought out by telemetric means. Torque in the port pinion shaft and intermediate shafting, together with the axial move- ‘ment of the gearcase and shafting, were recorded, as was the ‘radial displacement of the main wheel shaft in its bearings. ‘A static shaft alignment curve was also plotted. ‘The trials showed that wheel tooth bending stresses at the after end of the main wheel mesh were nearly double those measured forward. The repeated stresses varied cyclically over the least number (S) of main wheel revolutions which corres- ponded to a whole number of pinion revolutions (18) with a ‘maximum of 13°5 ton/in® at the port mesh, These indications ‘were confirmed by the variation in rim stresses, and later measure- ‘ments revealed some of the high frequency strains that have to be resisted by the shrink fit. ‘The variations in tooth bending could not be expected basically to increase the possiblity of tooth failure, though it is known that irregular cylinder pressures. can ‘considerably amplify torsional vibration stresses. The cyclic variation corres- pponded with the successive inter-meshing of the same combin: tion of teeth on pinion and wheel though whether this was attributable to skewing of the gear elements or 0 a local mesh- ing error causing an acceleration of the system could not be determined. Axial vibrations were, also present which would contribute to the general wear and tear of pinion and shaft assemblies. These factors, together with a 500 Hz tooth contact, frequency and the reduced damping of a system carried in ball and roller bearings, would have accounted for the fretting and slackening of pinion bearing securing arrangements found in these gears. This would, in turn, lead to changes in alignment and gear meshing and hence higher tooth stresses, and it was ‘concluded that all these factors taken together with the stress- raising defect found in the material had been sufficient to bring about failure. ‘This throws considerable responsibility on the gear manu facturer being able to accurately predict the running alignment of gears, having due regard to ship loading and movement of both gearbox and seatings in the hot full torque condition, This is especially so if pinion shaft alignment cannot be adjusted in the usual manner, ie. by lateral movement of bearing. Fractured teeth in the main reduction gear wheels of a 15000 hp motor cargo ship, leading to axial slip of the 23Marine Machinery Failures rim under the helical tooth forces, inspired a detailed investigation (Fig. 4). Alignment of the main wheel to the shafting was satisfactory and local malalignment between the main wheel and the two engine pinions was suspected. In particular, interest was directed toward the movement of the pinions under clutch plate forces. In this case, phased readings Fig. 4—Cracked main wheel rim ofthe tooth bending stresses showed the pinion to be “tumbling” lunder the action of a radial clutch force generated by a slight displacement ofthe clutch plates when engaging. Thiscould infact be stopped by de-clutching one pinion, and then reengaging when ‘both pinions were unningat the same speed, Measurements taken at the main wheel tooth roots showed superimposed stresses due to this movement. Re-design of the cluteh-pinion. bearing supports eliminated the clutch engagement effect and allowed {greater control of pinion alignment and hence more uniform tooth loading. In May 1971, the Society amended its rules to bring the tooth loading of épicyclic gears in line with those for parallel shaft gears. Up to that time the rules legislated only for parallel Shaft ‘gears and special consideration ‘was given to epcyclic designs, Experience with one design of epicyclic gear in landinstal- lations had shown that gears with tooth bending stresses of up to more than twice those in comparable marine pavallel shaft gears fave long and suocessul service. For the assessment of allowable foading, the planet or star wheel, whichis, of course, essentially an idler gear, has its allowable loading reduced by a factor Of 196 to take account of the reversal of foad on the tooth fanks, and the load per unit face width has been taken a the mean ‘of that for the sun and the annulus, which latter, being Softer, generally had a reduced face width compared with the sun pinion. Recently, the annulus face width has been increased to parity “There are currently two main types of steam turbine gearbox involving epieyelis. The original designs had a single primary epicyelic mesh for each turbine branch. Subsequenty, some installations witha large overall reduction ratio on account of low propeller rev/min were fitted with a. double epicyelic re- duction in the I-P. branch, the first reduction generally being Star gear. Most other epicyclic gear meshes ae of the planetary ‘ype. Examining the proposed tooth loadings for contact sltesoes, it would seem that under normal operating conditions With niirided teeth there should be no risk of pitting, even for the sun pinion~planet mesh. The only danger of surface damage in the annulus™planet wheel mesh would be if there was interference between the planet wheel tip and annulus gear root. This could normally only occur if the planet moved outward relative to the annulus and is thus partially dependent upon the flexibility of the annulus and its attachments. To minimize this possibilty, neeased tip rel hasbeen given to the planet get teeth, “Towards the middle of 1970 a series of failures commenced in four types of higher powered sets. In the first two types failures were confined tothe annulus rings of planetary type units and was mainly in the L.P. fst reduction gears. In the second two types fatigue breakages occurred in the teeth of the LP. first reduction star wheels. AS a result of the star ‘whee failures the Society imposed, in December 1970, a power, and consequently tooth loading’ restriction on these gears Pending the fiting of stronger H.P. 1 gears of coarser pitch. 214 In order to understand more of the behaviour of these sears under running conditions, the Society has co-operated With the turbine and gear manufacturers to take measurements of torque in the L-P. quill shaft and the outer coupling ring of the first reduction (planetary) gearing, telemetry techniques being employed. From these measurements it was concluded that the vibratory torques measured should not be responsible for the annulus ring failures, assuming equal torque to be transmitted by ‘each of the planet wheels. With this in mind, measurement of the relative movements of the various gear elements when under ‘operating conditions became important, the radial and torsional sliffness of the axial cross-section of the annulus rings having a direct bearing on the planet wheel tooth contact loads. Further work is planned in this field. Experiences with these gears isthe subject of a paper to be fGven 10 the Insitute in Apa, 197,50 will not be examined in ‘COUPLINGS AND cLUTcHES ‘Sometimes couplings are maligned as well as mal-aligned but it seems the inclusion of a flexible coupling in an engine~ ‘gearbox line gives excuse for laxity. A classic example concerned a dry cargo ship with a S-cylinder 6000 bhp main oil engine driving the line shafting directly at 154 rev/min through a rubber tyre type flexible coupling. During the maiden voyage there ‘was trouble with fuel pump suction pipes necessitating operating the engine on four cylinders for eleven hours. Shortly afterwards the forward tyre of the coupling ruptured. Torsionals were naturally suspected, but on checking the shafting considerable ‘misalignment. was’ found which necessitated completely. re- hocking the main engine and thrust block, well over half an inch being machined off the main engine forward chocks. Subsequent trials showed satisfactory torsionals at shaft speeds ‘over 50 rev/min on both four and five cylinders. ‘The importance of careful alignment of any shafting system cannot be over-emphasised, and particular attention should be paid to envisage what will happen in the “hot” full power ‘condition. Many problems arise where, for example, gearboxes rest with one end on a cofferdam and the other on a lubricating oil drain tank, and itis strongly advocated that alignment should bbe checked “hot” whenever possible, ‘This situation existed on two sister ships, each with two PPielstick Diesel engines driving a single screwshaft through electro-magnetic couplings and single reduction gears. Tt was found impossible to set the couplings in the “cold” condition, so as to keep the air gap within the maker's limits when the ‘couplings were energized. In fact, the necessary ‘clearances in the main engine and pedestal bearing rendered it difficult to set up these couplings in the first place. The situation was partially eased by stiffening the self-aligning pedestal bearing, but the Society’s final recommendation was the fitting of new bearings throughout with clearances such that the magnetic coupling air gaps could be held within the required limits. ‘A clutch failure leading to a train of other troubles occurred on a roll-onjroll-off container vessel having a 14000 bhp 8 cylinder non-reversing oil engine running at 375 rev/min and driving a 3-bladed c.p._ propeller through an_ hydraulically ‘operated friction clutch. TThe vessel was not to Lloyd's Register Class. See Fig. 5. ‘During manoeuvring two Diesel-driven alternators supplied. electrical power, while on passage power came entirely from the ‘main engine driven alternator. Automatic disengagement of the main engine-driven alternator from the bus bars when changing to the independent supply was incorporated within the electrical system and fed to a temporary black-out, itself overcome automatically ‘The clutch, propeller controls and main thrust bearing were housed within a common casing. To engage the clutch, pressure oil was admitted to spaces between the inner cases, and to disengage, this pressure was released and further pressure oil admitted to spaces at opposite ends of the inner cones. The history of the vessel revealed an inexplicable clutch seizure during the maiden voyage, followed by choking of the SJESEE EE, W7E "Dov, Operating Experience and Development Focal ofMarine Machinery Failures main lubricating ol filters, mainly with fine particles, and the ‘Whole oll charge contaminated, This trouble ecurred, when the Inno engine crankshaft was found to. haye moved forward Gamaging the locating bearing and the timing gear train. ‘The aftermost friction lining of \the clutch was worn, while the forward lining was in good condition. Mtr oe ore Perward nar cone rig pine ‘ROP sengsare} shart ‘ern fom ‘ewrnge espace ‘rain holes rte pace Fig, 5—Clutch arrangement ‘When changing from manoeuvring to passage alternators, the blackout was long enough to cause a marked drop in clutch oil engaging pressure, and automatic re-starting of the oil pump could bring a sudden application of full toad torque on the clutch. Since the design was such that any residue of drain oil jn the void space in the clutch was at this point thrown out ‘between the forward lining and the clutch shaft, it was apparent ‘that full load torque could be thrown suddenly on the aftermost lining after every blackout. Tt was apparent that a major re- appraisal both of clutch design and of the eles sstem was feat catastrophic failure, this time of an TLP. turbine primary pinion forward flange led to fire, totally gutting the engine Foom and fortunately, no one was seriously hurt, Between the time of building and the failure the flexible coupling had been renewed, At that time a small crack was found in the face of the HLP. pinion coupling flange which was stated to have started in the rim of the flange, and now passed down through a bolt hole toward the centre of the flange. In view of the owner's commitments and non-availability of spares, the flange was considered to remain efficient for a limited period. Failure finally occurred after a further 114 days steaming. From the outset it was clear that the engine room fire was consequential on the flexible coupling failures, which allowed ubricating oil to spray on the H.P. turbine. The remainder of the significant damage in way of the H.P. turbine and pinion ‘unit was confined to five components. See Fig. 6, At the instant of failure the watchieping engineer succeeded in closing the main stop valve, ‘The sequence of damage was clear, After the initial failure of the pinion shaft at its forward Fro. 6—Flexible coupling failure Trans. Mar.B., B72, Vol 8 flange, events would have moved rapidly. It seems likely that as failure of the pinion shaft occurred across the final holding area, ‘which was offset from the central axis ofthe pinion, the pinion flange turned out from the centte axis still restrained by the stiffness of the now distorting flexible coupling, It must be remembered that the fore and aft float of the pinion was re- Strained by it being a double helical gear, and the float of the HLP. rotor restrained by its thrust pads. Therefore, the pinion flange and coupling sleeve would be prevented from fying out until the rim of the flange had cleared the end of the pinion shaft, and this would account for this shaft end and the pinion flange back-face, being hammered. How long this took one cannot say, probably no more, than seconds, during which uch of the energy of the turbine, would have been absorbed in bling out the afer turbine bearing. ‘Pinion flange finally broke on upward swing, coming ‘out and carrying with it the flexible coupling cover. The coupl sleeve appeared to have worked itself out ofthe turbine coupling ring in two ot three revolutions, heavily distorting the teeth on its way. The H.P. rotor was undoubtedly still turning at this ‘moment but must have seized shortly after, causing the whirling ‘coupling flange and shaft to break. On opening the HP. turbine later, it was seen that at the moment of failure of the rotor shaft, it must have been nearly at standstill as the damage was Confined to a small part of the shrouding of the aft two rows of ‘Al the evidence at this stage of the investigation pointed to,an favard propagation of the crack in the ‘inion ange a being the origin of failure, “The metallurgical examination showed not one, but three distinct origins of failure. See Fig. 7: shaft forward flange 4) Through a point on the lip of the bore and on the face cof the flange, that had propagated round the shaft so as {o sever it completely fi) A second crack having an origin adjacent to a sharp ‘ep forming the rim of the flange; fit A Tongitudinal crack extending from the, bore surface of the central hole and penetrating the flange radially for & distance of 1-2in. This was a brittle fracture hhaving the appearance of a clink in the original forging, _most of which had been removed inthe boring operation, It was concluded thatthe initial crack propagation probably emanated from this large latent defect in the, bore, possibly Formed during heat treatment of the pinion forging. "The moral of this failure, would, with hindsight, be that things are not always. as they seem—a truism which can be applied many times in investigating machinery breakdowns. CCRANKSHAFTS tis disturbing to realize the incidence of crankshaft failure 215Marine Machinery Failures in the sixties, as a percentage of those at risk, was almost as {great as in the fifties. One of the reasons for this may be found in the development of the medium speed Diesel for marine use, ‘where the defect picture would be influenced by teething troubles ‘on this type of engine. Apart from this, we are still frced with, sharply radiused files, finely angled oil holes and similar stress raisers, all of which contribute to the reduction of the fatigue strength of the crankshaft. The highest incidence rates come, as might be expected, from “slipping” of built-up shafts, and ‘there is recent evidence to suggest that the factors of safety against slip for such shafts are much lower than the theoretical values. ‘A major breakdown concerned the starboard oil engine of a twin, 14-cylinder, V-engine installation, developing a total of 11480 bhp. The crankshaft failed without prior warning