The shaft-and-bearing assembly in a freon compressor was subjected to severe pounding and vibration after six years of service. After about one year of service, the compressor had been shut down to replace a bearing seal. One month before the shaft failed, a second seal failure occurred, requiring the collar, spacer sleeve, seal, roller bearing, and lock washer to be replaced. The shaft was made of 4140 steel, heat treated to a hardness of 20 to 26 HRC. The seal, bearing, and lock washer were commercial components. Investigation (visual inspection, 4.5x images, x-ray diffraction, hardness testing, and microscopic exam) supported the conclusion that shaft failure was initiated by fretting between the bearing race and the bearing surface on the shaft because of improper bearing installation. Once clearance was established between the bearing and the shaft, the shaft began pounding on the inner bearing race, causing final failure of the shaft surface. Recommendations included proper fitting of the shaft and bearing race to preventing movement of the bearing on the shaft. Also, the lock washer and locknut must be installed properly.

A steam accumulator, constructed with 10.3 mm thick SA515-70 steel heads and an 8 mm thick SA455A steel shell, ruptured after about three years of service. The accumulator was used in plastic molding operations. An extensive lack of weld penetration in this the head-to-shell girth weld was revealed by laboratory examination. Some misalignment of the head to the shell because of radial displacement of the shell and head centerlines was observed which was found to result in excessive clearances between the two parts and a slight difference in the thicknesses of the parts. Transgranular fracture with occasional secondary branching was revealed. It was interpreted by stress analysis that a small amount of misalignment added to lack of penetration increased the stresses to near the tensile strength of the material. The failure was judged to be a short-cycle high-stress notch-fatigue failure.

A 58.4 cm (23 in.) diam heavy-duty brake drum component of a cable-wound winch broke into two pieces during a shutdown period. Average service life of these drums was two weeks; none had failed by wear. The drums were sand cast from ductile iron. During haul-out, the cable on the cable drum drove the brake drum, and resistance was provided by brake bands applied to the outside surface of the brake drum. Friction during heavy service was sufficient to heat the brake drum, clutch mount, and disk to a red color. Examination of the assembly indicated that the brake drum would cool faster than its mounts and would contract onto them. Brittle fracture of the brake drum occurred as a result of thermal contraction of the drum web against the clutch mount and the disk. The ID of the drum web was enlarged sufficiently to allow for clearance between the web and the clutch mount and disk at a temperature differential of up to 555 deg C (1000 deg F). With the adoption of this procedure, brake drums failed by wear only.

A 230 mm (9 in.) diameter disk attrition mill was scheduled to grind 6.35 mm (0.25 in.) diameter quartz particles to a 0.075 mm (0.003 in.) diameter powder. Due to severe wear on the grinding plates, however, the unit was unable to complete the task of grinding the rock. The mill consisted of a heavy gray cast iron frame, a gravity feeder port, a runner, and a heavy-duty motor. The frame and gravity feeder weighed over 200 kg (440 lb) and, in some areas, was over 25 mm (1 in.) thick. To obtain the operating speed of 200 rpm, a gear system was used to transmit the torque from the 2-hp motor. The runner consisted of a 50 mm (2 in.) diameter shaft and two gray cast iron grinding plates. Investigation (visual inspection, historical review, photographs, model testing of new plates, chemical analysis, hardness testing, optical macrographs, and optical micrographs) supported the conclusion that the primary feed material was harder than the grinding plates, causing wear and eventual failure. Recommendations included reducing the clearance between the flutes and possible material changes.

An oil scavenge pump was found to have failed when a protective shear neck fractured during the start of a jet engine. Visual inspection revealed that the driven gear in one of the bearing compartments was frozen as was the corresponding drive gear. Spacer wear and thermal discoloration (particularly on the driven gear) were also observed. The gears were made from 32Cr-Mo-V13 steel, hardened and nitrided to 750 to 950 HV. Micrographic inspection of the gear teeth revealed microstructural changes that, in context, appear to be the result of friction heating. The spacers consist of Cu alloy (AMS4845) bushings force fit into AA2024-T3 Al alloy spacing elements. It was found that uncontrolled fit interference between the two components had led to Cu alloy overstress. Thermal cycling under operating conditions yielded the material. The dilation was directed inward to the shaft, however, because the bushing had only a few microns of clearance. The effect caused the oil to squeeze out, resulting in metal-to-metal contact, and ultimately failure.

Upon arrival at the erection site, an AISI type 316L stainless steel tank intended for storage of fast breeder test reactor coolant (liquid sodium) exhibited cracks on its shell at two of four shell/nozzle fillet-welded joint regions. The tank had been transported from the manufacturer to the erection site by road, a distance of about 800 km (500 mi). During transport, the nozzles were kept at an angle of 45 deg to the vertical because of low clearance heights in road tunnels. The two damaged joints were unsupported at their ends inside the vessel, unlike the two uncracked nozzles. Surface examination showed ratchet marks at the edges of the fracture surface, indicating that loading was of the rotating bending type. Electron fractography using the two-stage replica method revealed striation marks characteristic of fatigue fracture. The striations indicated that the cracks had advanced on many “mini-fronts,” also indicative of nonuniform loading such as rotating bending. It was recommended that a support be added at the inside end of the nozzles to rigidly connect with the shell. In addition to avoiding transport problems, this design modification would reduce fatigue loading that occurs in service due to vibration of the nozzles during filling and draining of the tank.

The drive wheel on a clutch-drive support assembly was slightly loose and caused clutch failures in service after 680,000 cycles. After failure, removal of the taper pin holding the drive wheel on the shaft was difficult, indicating that the pin was tight in the assembly. The taper pin was made of 1141 steel, the shaft 1117 steel, and the drive wheel 52100 steel. It was found that failure of the clutch-drive support assembly occurred as a result of fatigue fracture of the taper pin. A loose fit between the drive wheel and the shaft and between the drive wheel and the pin permitted movement that resulted in fatigue failure. Fretting of the pin and drive shaft was observed but did not appear to have contributed to the failure. To prevent reoccurrence, the assembly should be redesigned to include an interference fit between the shaft and the drive wheel. The drive wheel and the shaft should be taper reamed at assembly to ensure proper fit. In addition, receiving inspection should be more critical of the components and accept only those that meet specifications.

Plastic product failures are directly attributed to one of the following four reasons: omission of a critical performance requirement, improper materials specification, design error, and processing/manufacturing error. Therefore, product failures can be minimized or eliminated if all of these parameters are comprehensively examined during the design process. This article focuses on all of these factors, except processing-related failures, which are outside the design and engineering domain. It is dedicated to the identification and avoidance of common problems associated with the selection and designing of plastic parts. The article provides information on the material-related design criteria that depend on the applications, environmental conditions of use, and performance requirements. It discusses physical properties of plastics based on their relevance to real-world environmental conditions. The most-common design problems related to design considerations are also covered.

A mechanical part, which supports the moving part, is termed a mechanical bearing and can be classified into rolling (ball or roller) bearings and sliding bearings. This article discusses the failures of sliding bearings. It first describes the geometry of sliding bearings, next provides an overview of bearing materials, and then presents the various lubrication mechanisms: hydrostatic, hydrodynamic, boundary lubrication, elastohydrodynamic, and squeeze-film lubrication. The article describes the effect of debris and contaminant particles in bearings. The steps involved in failure analysis of sliding bearings are also covered. Finally, the article discusses wear-damage mechanisms from the standpoint of bearing design.

This article is dedicated to the fields of mechanical engineering and machine design. It also intends to give a nonexhaustive view of the preventive side of the failure analysis of rolling-element bearings (REBs) and of some of the developments in terms of materials and surface engineering. The article presents the nomenclature, numbering systems, and worldwide market of REBs as well as provides description of REBs as high-tech machine components. It discusses heat treatments, performance, and properties of bearing materials. The processes involved in the examination of failed bearings are also explained. Finally, the article discusses in detail the characteristics and prevention of the various types of failures of REBs: wear, fretting, corrosion, plastic flow, rolling-contact fatigue, and damage. The article includes an Appendix, which lists REB-related abbreviations, association websites, and ISO standards.

The failure of a high speed steel twist drill which caused injury to the user was investigated thoroughly to settle a legal suit. The drill was being used to remove a stud that broke in the vertical wall of a metalworking machine (upsetter) after drilling a pilot hole. The drill had shattered suddenly with a bang which caused a chip to be dislodged and cause the injury. A large nonmetallic inclusion parallel to the axis near the center of the drill was revealed in an unetched longitudinal section. Carbide bands in a martensitic matrix were indicated in an etched sample. It was concluded by the plaintiff's metallurgist that the failed drill was defective as the steel contained nonmetallic inclusions and carbide segregation which made it brittle. It was revealed by the defendant that the twist drill met all specifications of M1 high-speed steel and investigated several other drills without failure to prove that the failure was caused by use in excessive conditions. It was revealed by examination that the point of the broken drill was not the original point put on at manufacture but came from regrinding. Both technical and legal details have been discussed.

An LNG tanker experienced a fracture of the solid tail shaft, which is a section of the main drive shaft. The tail shaft was made of a forged low-carbon steel. In spite of two ultrasonic inspections, a large defect the size of a football in the center of the shaft was missed. During heat treating following forging, it was surmised that the defect led to the propagation of an internal brittle crack, or clink. A fatigue crack propagated from this origin to the outer surface of the shaft after about a year of service. Finally a last ligament of a few square inches held the shaft together and broke, leading to the separation of the shaft. The cause of failure was fatigue crack initiation and crack growth under reverse bending cyclic stresses. There was no indication that misalignment existed because there was no indication of fretting at the bolt holes in the flange at the end of the shaft. In the case of this shaft, a solution would have been to machine the core of the shaft to remove the brittle material or to use a tubular shaft.

While undergoing vibration testing, a type 347 stainless steel inlet header for a fuel-to-air heat exchanger cracked in the header tube adjacent to the weld bead between the tube and header duct. Investigation (visual inspection and liquid penetrant inspection) supported the conclusion that the crack in the header tube was the result of a stress concentration at the toe of the weld joining a doubler collar to the tube. The stress concentration was caused by undercutting from poor welding technique and an unfavorable joint design that did not permit a good fit-up. Recommendations included manufacturing the doubler collar so that it could be placed in intimate contact with the header duct, and a revised weld procedure was recommended to result in a smaller, controlled, homogeneous weld joint with less distortion.

Hydraulic cylinders on three identical presses failed in a similar manner after approximately ten years' service life. The cylinder was a steel casting having a carbon content of the order of 0.3 to 0.4%. During machining of the internal surfaces, a sharp corner had been left at the junction of the head with the shell. From this stress raiser a fatigue crack had developed around the entire circumference of the cylinder to give a smooth crack of annular form. The use of a flat end to the cylinder, therefore, resulted in excessive stresses being introduced at the junction of the end with the cylinder.

A pipe in a chip conveyor cracked at the toe of an exterior fillet weld connecting a flange to the pipe. The chip conveyor consisted of several spool sections. Each section was made up of a length of low-alloy steel pipe and two flanges, which were welded to each end. The composition specified for the pipe steel was 0.25C-0.98Mn-3.52Ni-1.34Cr-0.24Mo, which approximates a 9300 steel with high molybdenum. Investigation supported the conclusion that the conveyor pipe failed by brittle fracture, which was attributed to the stresses induced in forcing the circular flange over the elliptical section of the pipe. The toe of the weld and the adjacent undercut were stress raisers that determined the point of major crack origin. Under residual stress, the internal point of incomplete fusion also initiated additional cracks. Recommendations included ensuring a proper fit between an elliptical flange and pipe end to eliminate the cracking.

A 9310 steel gear was found to be defective after a period of engine service. A linear crack approximately was discovered by routine magnetic-particle inspection of an electron beam welded joint that attached a hollow stub shaft to the web of the gear. The welding procedure had a cosmetic weld pass on top of the initial full-penetration weld. There were no other known service failures of gears were welded by this method. One zone of the welded joint showed incomplete fusion, surrounded by two zones containing fatigue beach marks This indicated that the incomplete-fusion zone was the site at which primary fracture originated. The possible causes of incomplete-fusion include localized magnetic deflection of the electron beam, a momentary arc-out of the electron beam, and eccentricity in the small weld diam. The failure was attributed to fatigue originating at the local unfused interface of the electron beam weld, which had been the result of a deviation in the welding procedure. Examination of the possible causes of failure gave no evidence that a recurrence of the defect had ever occurred. Thus, there was no basis on which to recommend a change in design, material, or welding procedure.