Edited by Martha K. Raymond
Designing a joint and selecting fasteners are more of a challenge than most people realize. First and foremost, the engineering physics of joints and fasteners is one of the least intuitive aspects of engineering.
The reason is that both the joint and its fasteners are elastic bodies that compress and elongate in complex ways quite crucial from an engineering standpoint. This applies regardless of how large or small a structure is. Other important considerations are the type of materials being joined, environmental conditions, the style of the structure, the required service life of a joint, and the tooling and methods to be used in assembly.
Analyzing the working forces acting on the joint are part of a design. For example, the types of loading may include shear, bending, tension, or fatigue. These loads can have a significant impact on joint failure. The unforeseen loading conditions that can lead to joint failure are fatigue failure of the bolt, insufficient preload, vibration loosening of the bolt, yielding of the joint material under the bolt or nut, and tensile overload of the bolt. With so many possibilities for failure combined with options in joint design, some general guidelines for structural design can be applied and can lead to a strong, reliable joint.
One of the first considerations is installation control which also encompasses joint design and fastener selection. The longer the joint, the greater the total bolt elongation that’s required to produce the proper preload. Typically, the longer the joint, the less potential there is for loss of preload. Another factor to consider is joint material. A material that is relatively stiff compared to the bolt material will compress less. Therefore, stiff material provides a joint that is less sensitive to preload loss as a result of relaxation, brinelling, or loosening.
Next, consider the thread stripping strength of the joint material. The material in which the threads will be tapped or the nut used must allow sufficient engagement length to carry the load. Ideally, the length of thread engagement should be sufficient to break the fastener in tension. When a nut is used, the wall thickness of the nut as well as its length must be considered.
Another step in designing a joint is to calculate or evaluate the tension loads that the bolt and joint will experience. The bolt size and the quantity necessary to carry the expected load, along with the safety factor, must also be determined.
While many applications apply a tensile load to the fastener, some apply a shear load. This type of loading is applied perpendicular to the fastener axis. Shear loading may be single, double, or multiple. Allowable shear strength as listed in handbooks or product data is listed in pounds per square inch, which is the shear load in pounds divided by the cross sectional area in square inches. In a comparison of single and double shear strength, single shear strength is exactly one-half the double shear value.
However, materials have a relationship between the tensile and shear strength. For example, in alloy steel, the shear strength is 60% of its tensile strength. On the other hand, corrosion-resistant steels, such as 300-Series stainless steels, have a lower tensile/shear relationship, usually 50 to 55%.
Another design consideration is application temperature. For elevated temperature, standard alloy steels are useful to about 550 to 600°F. However, if plating is used, the maximum temperature may be less. For example, cadmium should not be used over 450°F. Austenitic stainless steels, the 300 Series, may be useful to 800°F. They can maintain strength above 800°F, but the surface will begin to oxidize. The operating environment also plays a role in design. A plating may be selected for mild atmospheres or salts. If plating is unsatisfactory, a corrosion-resistant fastener may be specified. The proper selection will be based upon the severity of the corrosive environment.
Another factor is fatigue strength. The nature of fatigue is such that under repetitive cyclic loading or stressing, failure can occur at strength levels well below the rated tensile strength. Data for fatigue strength usually comes from a S/N graph, where S is stress plotted on the X axis and N is the number of cycles plotted on the Y axis. On this type of curve, the high load to low load ratio must be shown, which is usually R = 0.1, where the low load in all tests will be 10% of the high load.
The R value must change to account for preload. Increasing the R to 0.2, 0.3 or higher will change the curve shape. At some point in this curve, the number of cycles will reach 10 million, which is defined as the “endurance limit” or the stress at which infinite life may be expected. Nevertheless, the S/N curve is based on theory and doesn’t provide the necessary information to determine the performance of a specific fastener. In application, the preload should be higher than any of the preloads on the S/N curve. Therefore, a modified Goodman Diagram or Haigh Soderberg Curve can be used to show what fatigue performance can be expected when parts are properly preloaded.
The preload of a fastener can be calculated by several methods. One way is by elongation. For example, the modulus for steel of 30,000,000 psi means that a fastener will elongate 0.001 in./in. of length for every 30,000 psi applied in stress. Therefore, if 90,000 psi is the desired preload, the bolt must be stretched to 0.003 in. for every inch of its length in the joint.
This preload method is accurate but requires properly prepared bolt ends and precise measurements. In addition, direct measurements are only possible where both fastener ends are available for measurement after installation.
A second method of preload is turnof- the-nut. This process also uses change in bolt length. In theory, one bolt revolution, 360° rotation, should increase the bolt length by the thread pitch. There are at least two variables, however, which influence this relationship. One is that a joint needs to be snug to measure bolt elongation, but the snugging produces a large variation in preload. Secondly, the joint undergoes compression, so the relative stiffness of the joint and bolt effects the final load.
A third way to determine preload is through strain on the joint. Because stress versus strain is a constant relationship for any given material, that relationship can be used in the same way as elongation measurements. The strain can be detected from strain gages applied directly to the outside and inside of the bolt. The output from these gages need instrumentation to convert the gage’s electrical measurement to strain indicators. The strain measurement method is, however, expensive and not always practical.
By far, the most popular method of preloading is by torque. Fastener manufacturers usually have recommended seating torques for each size and type of fastener material. The only requirement is that the proper size torque wrench and correct torque amount are used.
Because torque applied to a fastener must overcome all friction before any loading takes place, the amount of friction present is important. In a standard unlubricated assembly, the friction to be overcome is in the head bearing area and the thread-to-thread friction. Approximately 50% of the torque applied will be used to overcome head-bearing friction, and approximately 35% is used to overcome thread friction. Consequently, 85% of torque is overcoming friction and only 15% is available to produce bolt load. If the interfaces are lubricated, for example with materials such as cadmium plate, molydenum disulfide, or antiseize compounds, friction is reduced and greater preload can be produced with the same torque.
The data between torque and tension is typically plotted on a torquetension curve. However, the change in the coefficient of friction for different conditions can significantly affect the slope of the curve. If this is not taken into consideration, the proper torque specified for a plain unlubricated bolt may be sufficient to yield or break a lubricated fastener.
Another consideration for determining torque is thread pitch. It needs to be included where a given stress is to be applied. The cross-sectional area used for stress calculations is the thread tensile-stress area, and it differs for coarse and fine threads. Torque recommendations to reach the same stress are slightly higher for fine threads than they are for coarse threads.
There are major differences between coarse and fine threaded fasteners. Coarse threads require fewer turns and reduce cross threading. They have higher thread-stripping strength per given length and are more common in industrial fasteners. However, they exhibit lower tensile strength, reduced vibrational resistance, and are capable of only coarse adjustment.
On the other hand, fine threads provide higher tensile strength, greater vibrational resistance, and allow finer adjustment. Their disadvantages include easy cross threading, threads easily damaged through handling, more critical tap drill size, and slightly lower thread-stripping strength.
In addition to the joint design factors, other considerations are important to properly use high-strength fasteners. One is that the mating nut height is properly chosen to guarantee adequate thread engagement. Also, the minimum engagement length recommended in a tapped hole depends on the strength of the material, but in all cases it should be adequate to prevent stripping. Another consideration is to specify a nut with the proper strength level. The nut and bolt should be selected as a system. And yet another step is to specify compatible mating female threads. A possible combination includes 2B tapped holes or 3B nuts.
Another problem is corrosion. In general, corrosion penetrates the joint and not just the bolt alone. This can be a matter of galvanic action between dissimilar metals. Corrosion of the fastener material surrounding the bolt head or nut can be critical with high strength bolting. To prevent corrosion, joint material should be compatible, and coatings can be used to protect dissimilar materials.