The vast majority of threaded fasteners retain their preload because of friction present on the thread and head contact surfaces that resists self-loosening. It was thought that once relative motion occurs between the threaded surfaces and other contact surfaces of the clamped parts (because of a tangential external force being applied to the joint), the bolted connection would become free of friction in a circumferential direction. More recent studies have shown that under conditions of transverse slip there is a small, but measureable, force resisting 
circumferential movement.
In the above diagram, the red arrow shows the joint surface moving under the nut. When slip occurs at the thread interface as well as the nut face, the frictional resisting forces in the circumferential direction reduce to a very small value. The Jost Effect is the name given to the reduction in the frictional resistance that occurs in a direction different to that inwhich slip is occurring. This effect is used in many applications including the removal of corks from bottles. If the cork is first rotated the force needed to pull the cork from the bottle is significantly reduced. Several machines and applications use indirectly the Jost Effect, for example floor polishing machines, the machine being easier to move when the polishing disk is rotating. It is also the fundamental reason why threaded fasteners experience self-loosening. Frictional resistance is first overcome in the transverse direction by slip occurring on the joint resulting in the frictional resistance in the circumferential direction reducing to a small value. The torque acting on the fastener in the loosening direction (as a result of its preload) that when coupled with the Jost Effect results in self-loosening occurring. This is illustrated in the diagram.
The term is named after the Institute that completed research into this effect, the Jost Institute of Tribotechnology at the
Strength of Threaded Fasteners
Fastener failure on a product can have potentially disastrous consequences. In an attempt to ensure that such consequences do not occur, rigorous and extensive testing of a product is frequently completed. However in many applications, extensive testing is neither practical nor economic. In such instances, the Engineer usually relies upon analytical analysis together with his experience and judgment to ensure that failure does not occur.
Failure of a threaded fastener generally occurs in one of three modes. Failure through the shank or threaded section of the fastener, thread stripping of the external thread, or thirdly, thread stripping of the internally threaded member. Considering each in turn:
Failure through the male thread or thread shank
The majority of fastener failures occur with fracture through the male thread. Under static loads, the strength of the thread is determined by the stress area. This is based upon the mean diameter of the minor and effective thread diameters. Engineering handbooks have, typically, tables of stress areas for various thread sizes.
When a bolt is tightened the shank sustains a direct stress, due to the elongation strain, together with a torsional stress, due to the torque acting on the threads. Most tables of bolt tightening torques ignore the torsional stress and assume a direct stress in the threads of some proportion of the bolts yield stress, usually 75%. For high frictional conditions the magnitude of the torsional stress can be such that when combined with the direct stress, an equivalent stress over yield can result, leading to failure. A more consistent approach is to determine the magnitude of the direct stress which, when combined with the torsional, this will give an equivalent stress of some proportion of yield. The proportion commonly used with this approach is 90%. The computer program TORQUE provides state of the art analysis for the torque tightening of threaded fasteners.
High performance bolts are often designed so that the plain shank is smaller than the stress diameter of the thread. This is done so that the stretch that occurs under the preload induced from the tightening process is maximised. With this type of bolt, failure, if overtightened, will occur in the plain shank region as shown in the photograph.
Thread Stripping
Thread stripping can be a problem in many designs where tapped holes are required in low tensile material. In general terms, thread stripping of both the internal and external threads must be avoided if a reliable design is to be achieved. If the bolt breaks on tightening, it is obvious that a replacement is required. Thread stripping tends to be gradual in nature. If the thread stripping mode can occur, assemblies may enter into service which are partially failed, this may have disastrous consequences.
The photograph above is from a scanning electron microscope showing a bolt thread about to strip. The joint surface interface was at the right hand side, you can see from the image that the first thread has the greatest distortion. The thread stripping mechanism is complex and involves thread bending (that occurs under the high loads) and nut dilation (which results in the shear plane moving).
To precisely predict the force and mode of failure of a threaded assembly demands consideration of a large number of factors. Thread stripping is a complex phenomenon. The following factors all have an important effect on the stripping strength of a thread:
1. The variation in the dimensions of the thread, (such as major, pitch and minor diameters) has a significant effect on both internal and external threads stripping strength.
2. Tensile and shear strength variations in the material for both the internal and external threads.
3. The effect of radial displacement of the nut or tapped component (generally known as nut dilation) in reducing the shear strength of the threads. The tensile force in the fastener acts on the threads and a wedging action generates a radial displacement which reduces thread strength.
4. The effect of bending of the threads, caused by the action of the fastener's tensile force acting on the vee threads, resulting in a wedging action that decreases the shear area of the threads.
5. The effect which production variations in the threaded assembly, such as slight hole taper or bellmouthing, can have on thread strength.
The strength of a nut or bolt thread cannot be viewed in isolation without considering the inter-dependence which both elements have on the strength of the assembly. One of the problems in predicting thread stripping strength is that, without considering such effects as thread bending, nut dilation or bellmouthing, an optimistic result occurs. The actual stripping strength being lower than the calculated. The program FASTENER allows a state of the art analysis to be performed to determine the stripping strength of a threaded assembly.
Bellmouthing Effects
A complicating factor which can occur when a drilled hole is tapped, is bellmouthing. This is a slight taper on the hole which is usually encountered on most drilled holes to some degree. This taper extends normally for about half the diameter from the start of the hole. The cause of this tapering is torsional and transverse flexibility of the drill together with instability of the drill point during entry into the material. Bellmouthing can be minimised by the use of close fitting, well aligned and rigid drill bushes together with accurate drill sharpening.
Holes exhibiting bell mouthing will, when tapped, experience a variable thread height along the length of the hole. This variation can be significant on short lengths of engagement and fine pitches. The net effect of bellmouthing is to reduce the shear area of the external threads. The finer the thread the more pronounced is the effect of bellmouthing.
Influence of tap-drill size on thread strength
In tapped holes, the thread height is dictated by the diameter of the tapping drill. To reduce the risk of failure, the Design Engineer is often cautious and specifies high percentages of thread height in tapped holes. From a production standpoint these higher percentages of thread height result in higher tapping torques, increased tap breakages and, as such, are not favored. For short lengths of thread engagement, the minor diameter size - resulting from the tapping drill - has a significant effect on assembly strength. Studies have shown that for threaded assemblies of usual proportions, tap-drill size is relatively unimportant so long as the percentage of thread height is greater than 60%. Tapping costs are likely to be lower if the lowest possible thread height is used.
The effect of a low proportion of thread height is to reduce the shear area of the external thread, this is illustrated in figure 1. For very low thread heights, the shear plane through the threads need not be parallel to the thread axis, this is illustrated in figure 2. Such failure modes are difficult to predict and can be easily eliminated by maintaining a reasonable percentage thread height.
Nut Dilation
The tensile force present in the fastener during tightening acts on the vee threads to produce a wedging action which results in a radial displacement. This radial displacement is generally known as nut dilation and occurs in threaded bosses as well as conventional nuts. Theoretical and practical studies of this phenomenon indicate that the top face of the nut contracts in a radial direction while its bearing surface expands. The net effect of this dilation is to reduce the shear area of both the internal and external thread.
The stripping strength of an assembly can be improved by increasing the width across flats of the nut, or boss diameter, up to about 1.9 times the nominal thread diameter. This increases the stiffness locally around the internal thread and reduces radial expansion.
Thread Bending
The tensile force present in the fastener during the tightening process results in a degree of thread bending between internal and external threads. Thread bending reduces the shear area of both internal and external threads. A dominating factor controlling the degree of thread bending is the ratio between the strength of internal and external threads. The strength ratio is the ratio between the force necessary to cause the nut thread to strip, divided by the force required to cause the bolt thread to strip.
Leakage From Joints Containing Gaskets
Introduction
Incidents of fluid leaking from a joint containing a gasket is one which many Engineers have encountered. The inclusion of a gasket in a joint to prevent leakage is only effective as long as there is sufficient clamp force generated by the bolts to allow effective sealing. For joints whose design are not covered by standards, establishing the number and distribution of bolts which must be used to ensure effective sealing can be problematical. The complexities of gaskets are such that empirical methods have been developed to approximate their performance and operating characteristics. These methods involve the use of empirically derived factors to approximate the clamp force required to allow the gasket to seal effectively.
There are two practical factors which are commonly used in joint design involving gaskets. The m or maintenance factor is used to establish the clamp force required to ensure an effective gasket seal when the joint is subjected to internal pressure. The y or yield factor is used to determine the clamp force required to be applied to a gasket to ensure that it seats properly to provide a seal. Further explanation, of how these factors are used to establish the bolt clamp force needed for the gasket to seal effectively, is presented in the boxed caption.
It must be pointed out that besides calculation, a large degree of experience is required to ensure a leak free joint. Gasket manufacturers sometimes vary the m and y factors to suit the medium which the gasket is sealing. This is based upon the experience that gaseous media is generally more difficult to seal than fluids. Increasing the factors in this way allows the approach to be valid where overwise it would fail to yield sound results. Consultation with gasket manufacturers is to be encouraged at the joint design stage to ensure that account is taken of all relevant factors.
Another common problem is inadequate bolt spacing. In many joints the bolt spacing is dictated by the gasket pressure mid-way between bolts. If insufficient pressure is applied to the gasket in such regions, leakage can result. Research into this subject indicates that clamping pressure, in joints without gaskets, quickly decays away from the bolt. In such joints zero pressure occurs between 1.8 and 3.6 times the bolt radius, the actual value being dependent upon the joint details. With a gasket the pressure decays more slowly, however, to overcome this problem many design codes dictate a maximum bolt spacing (which varies with bolt size).
Local crushing of the gasket can occur if the clamp force generated by the bolt is excessive for a particular gasket material. Special pressure sensitive film (such as
Ideally all the bolts of the joint should be tightened simultaneously, especially when gaskets are being used. If this is not achievable, a tightening sequence should be specified. A poorly specified tightening sequence will result in uneven load distribution in the joint increasing the chances of failure occurring. (Without using numerical techniques such as the finite element method and making assumptions about the unevenness of the joint faces, it is not possible to quantify the magnitude of this irregular load distribution.) Sound tightening sequences, based upon experience, have been established. If the bolts are in a circular pattern, a cris-cross tightening sequence is usually specified. For non-circular bolt patterns, a spiral pattern starting at the middle has been found to be effective.
Creep or relaxation of the gasket material can be a practical problem. This is a particular consideration in joints subjected to temperatures greater than 100 °C. The amount of relaxation which can occur with gaskets is usually far greater than that which occurs in joints which do not contain such compliant materials. Because of the magnitude of bolt clamp force loss which can occur with gaskets, frequently a re-tightening schedule is specified 24 hours or more after initial tightening. The magnitude of the clamp force reduction due to gasket creep can be of such magnitude that is not feasible to design the bolts for this loss. To overcome this problem, re-tightening of the bolts can be specified after a period of time following initial tightening, frequently 24 hours. Such a schedule may involve re-tightening the fasteners on a regular basis to overcome the problems of relaxation.
The most prevalent controlled method of tightening bolted joints containing gaskets is by tightening so that a specified torque is achieved. This method is generally known as torque control. Without experimentally obtained torque values or analytical tools, specifying the correct torque which should be used in an installation, can be problematical. Leaks from many joints are directly attributable to a poor torque specification. A deficient tightening torque leads directly to an inadequate clamp force. This may well be insufficient to achieve gasket sealing.
By a wider awareness of the potential problems involved when designing joints containing gaskets, and by an appreciation by Engineers of the importance of bolt clamp force, will gasket leakage problems be prevented. In order to assist the Engineer in the specification of the correct tightening torque, Bolt Science has developed a computer program which can account for the relevant factors involved. Using the assistance of on-line help screens and an in-built database, the Engineer can determine the tightening torque and the resulting clamp force for both metric and imperial threaded fasteners. The effect of changes in the fastener design and the prevalent frictional conditions can quickly be established.
The Gasket Factor m
If the joint containing a gasket is to seal an internal pressure then there must always be a sufficient clamp force applied to the gasket, to ensure that a leak free joint is achieved. The m (maintenance or multiple) factor, is the factor that provides the additional clamp force capability in the joint's fasteners to maintain sealing pressure on a gasket after an internal pressure is applied to the joint. It is a multiple of the internal pressure; the ratio of gasket contact pressure to contained pressure. Values of the m factor for various gasket materials can be obtained from pressure vessel and other similar national and international standards. Values for the m factor can vary from 0.5 for rubber to 3.5 for asbestos (values vary with thickness and composition). To provide an estimate of the minimum bolt clamping force required for joints subjected to an internal pressure, the following formula may be used with caution:
Fk = (Fh + P x m x A) / n
where:
Fk = Minimum required clamp force from each bolt in the joint
Fh = Hydrostatic end force acting on the joint
P = Internal pressure acting on the joint
m = Gasket factor
A = Total area of gasket based upon using an effective gasket width (this value is covered in standards such as BS 5500)
n = Number of bolts in the joint
The formula assumes that the same size and grade of bolt is used in the joint and that the distribution of bolts in the joint is sound. To establish what bolt pattern does give a sound joint is largely down to experience.
The Gasket Factor y
Before a leak free joint can be obtained, it is necessary to seat the gasket properly by applying a minimum initial pressure (under atmospheric temperature conditions without the presence of internal pressure). This design seating stress has been given the term y (yield) factor, and it is the stress required to deform the gasket into the irregularities of the joint surface. It is governed by the compressibility of the gasket material. Values of the y factor for various gasket materials are given in the BS 5500 standard. To determine the minimum clamp force to meet this requirement, the following formula can be used:
Fk = A x y/n where y is the gasket factor and other terms are as given before.
Conclusions
The minimum clamp force required from the bolts, to ensure that the gasket seals effectively, is the maximum of the two values determined using these two factors. Implicit in this method is that the bolt spacing and the rigidity of the joint flanges are such, that problems will not occur because of deficiencies in these areas. This is largely down to experience for those joints which are not covered by existing design codes.
Applying state of the art analytical analysis to prevent gasket leakage can be complicated. To assist the Engineer in overcoming the problems associated with the use of threaded fasteners and bolted joints, Bolt Science has developed a number of computer programs. These programs are designed to be easy to use so that an engineer without detailed knowledge in this field can solve problems related to this subject.
THE JOINT DECOMPRESSION POINT
Why Bolt Preload is Important!
The most common reason why bolted joints fail is due to the bolt failing to provide sufficient preload to prevent the external applied forces overcoming the clamp force acting between the joint faces.
The slide show presented below illustrates the joint decompression point. This is when the clamp force acting between the joint faces, that has been provided by the bolt's preload, has been reduced to zero by the applied forces.
Once the joint faces have separated the bolt will be subjected to bending forces and the joint faces to fretting. This will lead to a loss of preload and the bolt subsequently failing by fatigue or other mechanism. This is way the decompression point is taken as a design failure criterion.
The importance of having a high preload can be illustrated by using the decompression point. The image below illustrates this point:
As can be seen that a higher preload raises the decompression point. For this reason it is better to tighten a fastener up to close to its limit rather than only partially tightening it.
Tests have shown that the elastic interaction between bolts in a joint can have a significant effect on the preload (a reduction of 35%). If a gasket is present between the joint surfaces the effect can be even more pronounced.
TIGHTENING SEQUENCES
The Appropriate Sequence to Tighten a Joint
Because in the vast majority of bolted all bolts are not tightened simultaneously, the effect of tightening one bolt in the group as an effect on the preload in other previously tightened bolts in the group. Such effects are called elastic interactions or sometimes bolt crosstalk. The mechanism that causes this is illustrated in the diagram below.
The outer two bolts have been tightened compressing the joint under the bolts. The middle bolt is subsequently tightened compressing the joint directly under the bolt but also compressing the joint slightly under the two other bolts leading to a loss of preload in these bolts. Presented below are two examples of tightening sequences that have been shown to result in minimising bolt preload variations due to elastic interactions, that will minimise the preload scatter within a joint. If the joint is critical it would be wise to consider specifying a multiple pass tightening sequence. With such a sequence, each bolt is tightened more than once so as to reduce the preload reduction caused by the tightening of the other bolts in the joint.
TIGHTENING THE NUT OR THE BOLT?
The question is often asked as to whether the nut or the bolt head should be tightened. The answer depends upon which tightening process is being used. For torque controlled tightening whether the nut is tightened and the bolt head held, or the bolt head tightened and the nut held, can be of importance.
The general objective from a tightening process is to achieve a consistent bolt preload. Controlling the torque during tightening and completing subsequent inspection checks to ensure that the specified torque is being achieved, are common ways that this objective is implemented.
When applied torque and the resulting tension (preload) in the bolt are measured during tightening and plotted on a graph, there is a linear relationship between the torque and the tension. The bolt tension is directly dependent, and proportional to, the applied torque. This is illustrated by the graph, which is based upon experimental results, that is shown in the diagram above. From such test results it is possible to establish the appropriate torque for a required bolt preload that may be required.
One of the disadvantages of using torque control is that there can be a significant variation in the bolt preload achieved for a given torque value. There are several reasons for this e.g. inaccuracy in applying the torque, dimensional variations of the thread and hole size variation amoungst others. However, the dominant factor is usually due to the frictional variation that is present between the contact surfaces that are being rotated.
From tests, it is known that approximately 50% of the tightening torque is dissipated in overcoming friction under the bolt head or the nut face (whichever is the face that is rotated). Typically only 10% to 15% of the overall torque is actually used to tighten the bolt, the rest is used to overcome friction in the threads and on the contact face that is being rotated (nut face or bolt head). This is illustrated in the piechart shown above. Relatively small changes in the nut face friction can have a significant effect on the bolt preload. As more torque is perhaps needed to overcome friction, less remains for the bolt extension and hence as the effect of adversely reducing the preload. If the friction under the nut face is reduced, then, for a given torque, the bolt preload will be increased.
The diagram shown at the side is perhaps the most common situation where the top and bottom plates of the joint are made from the same material, have the same finish and the hole size is the same through both of the plates. For such a joint, when the nut face and bolt head sizes have the same diameter and finish, it will not matter whether the bolt head or the nut is tightened. Some people believe that by tightening the bolt head rather than the nut it will affect the torsion in bolt shank. The torsion in the shank of the bolt depends upon the thread friction torque. For a given finish condition, the thread friction has some scatter associated with it, but will not depend on whether the nut or the bolt head is tightened. If the thread friction torque remains the same, the torsion in the shank will be the same irrespective of whether the bolt head or the nut is tightened.
The diagram at the side shows the situation when the plates comprising the joint are different 
materials (such as one being steel and the other aluminium) or have different finishes (such as one plate being galvanised and the other painted). In such situations, it will, in general, be important as to whether the bolt head or nut is tightened. The reason is that each face will have a different friction coefficient. If the tightened torque was determined either by testing or by looking up the friction characteristics of the surface, say based upon the nut face, then it is probable that the head face would have a different friction coefficient. If it had a lower friction value then the preload would be increased if the bolt head was tightened. In the extreme case, if the frictional differences were large, bolt breakage could 
occur.
The diagram at the side illustrates the case when the clearance hole in the top plate differs from that used in the bottom plate. Such situations are relatively common. There is an effective friction radius on the part that is rotated (nut or bolt head) that is usually taken as the mean of the clearance hole and outer bearing face radii. Because this radii would be greater for the bolt head than the nut in the situation shown, less bolt preload would result by tightening the bolt head rather than the nut, other factors such as friction being the same. Hence another example of a situation as to whether the nut or bolt head is tightened.
The drawing at the side the case when there are style and dimensional differences between the bolt head and the nut. The effect is similar to that which happens in the previous case. Differences in the friction radii between the bolt head and the nut-washer interface result in the preload being affected by which item is tightened. In the case shown here, there would probably also be differences between the friction coefficient that is present when the nut is tightened on the washer and the bolt head onto the joint. This would increase the variability still further.
Washers are occasionally used as a means of minimising frictional scatter besides the common reason of reducing the bearing stress on the joint face. The friction condition between the washer and nut face can be reasonably well defined and controlled, more so than the joint surfaces usually can. By controlling the friction, the preload can be more reliably achieved. To do this consistently, a close fit is needed on the inside diameter of the washer. One way in which this can be achieved is by the use of a SEMS unit (in which a washer is held captive on the bolt shank). The same can be achieved by using a KEPS unit (a washer being held captive on a nut).
So in general, when using torque control, tightening the bolt by rotating the bolt head or the 
nut can matter. It is good practice to specify which part should be tightened so that the bolt preload variation is minimised.
To assist the Engineer in overcoming the problems associated with the use of threaded fasteners and bolted 
joints, Bolt Science has developed a number of computer programs. These programs are designed to be easy to use so that an engineer without detailed knowledge in this field can solve problems related to this subject.
Tightening the Nut or the Bolt?
The question is often asked as to whether the nut or the bolt head should be tightened. The answer depends upon which tightening process is being used. For torque controlled tightening whether the nut is tightened and the bolt head held, or the bolt head tightened and the nut held, can be of importance.
The general objective from a tightening process is to achieve a consistent bolt preload. Controlling the torque during tightening and completing subsequent inspection checks to ensure that the specified torque is being achieved, are common ways that this objective is implemented.
When applied torque and the resulting tension (preload) in the bolt are measured during tightening and plotted on a graph, there is a linear relationship between the torque and the tension. The bolt tension is directly dependent, and proportional to, the applied torque. This is illustrated by the graph, which is based upon experimental results, that is shown in the diagram above. From such test results it is possible to establish the appropriate torque for a required bolt preload that may be required.
One of the disadvantages of using torque control is that there can be a significant variation in the bolt preload achieved for a given torque value. There are several reasons for this e.g. inaccuracy in applying the torque, dimensional variations of the thread and hole size variation amoungst others. However, the dominant factor is usually due to the frictional variation that is present between the contact surfaces that are being rotated.
From tests, it is known that approximately 50% of the tightening torque is dissipated in overcoming friction under the bolt head or the nut face (whichever is the face that is rotated). Typically only 10% to 15% of the overall torque is actually used to tighten the bolt, the rest is used to overcome friction in the threads and on the contact face that is being rotated (nut face or bolt head). This is illustrated in the piechart shown above. Relatively small changes in the nut face friction can have a significant effect on the bolt preload. As more torque is perhaps needed to overcome friction, less remains for the bolt extension and hence as the effect of adversely reducing the preload. If the friction under the nut face is reduced, then, for a given torque, the bolt preload will be increased.
The diagram shown at the side is perhaps the most common situation where the top and bottom plates of the joint are made from the same material, have the same finish and the hole size is the same through both of the plates. For such a joint, when the nut face and bolt head sizes have the same diameter and finish, it will not matter whether the bolt head or the nut is tightened. Some people believe that by tightening the bolt head rather than the nut it will affect the torsion in bolt shank. The torsion in the shank of the bolt depends upon the thread friction torque. For a given finish condition, the thread friction has some scatter associated with it, but will not depend on whether the nut or the bolt head is tightened. If the thread friction torque remains the same, the torsion in the shank will be the same irrespective of whether the bolt head or the nut is tightened.
The diagram at the side shows the situation when the plates comprising the joint are different materials (such as one being steel and the other aluminium) or have different finishes (such as one plate being galvanised and the other painted). In such situations, it will, in general, be important as to whether the bolt head or nut is tightened. The reason is that each face will have a different friction coefficient. If the tightened torque was determined either by testing or by looking up the friction characteristics of the surface, say based upon the nut face, then it is probable that the head face would have a different friction coefficient. If it had a lower friction value then the preload would be increased if the bolt head was tightened. In the extreme case, if the frictional differences were large, bolt breakage could occur.
The diagram at the side illustrates the case when the clearance hole in the top plate differs from that used in the bottom plate. Such situations are relatively common. There is an effective friction radius on the part that is rotated (nut or bolt head) that is usually taken as the mean of the clearance hole and outer bearing face radii. Because this radii would be greater for the bolt head than the nut in the situation shown, less bolt preload would result by tightening the bolt head rather than the nut, other factors such as friction being the same. Hence another example of a situation as to whether the nut or bolt head is tightened.
The drawing at the side the case when there are style and dimensional differences between the bolt head and the nut. The effect is similar to that which happens in the previous case. Differences in the friction radii between the bolt head and the nut-washer interface result in the preload being affected by which item is tightened. In the case shown here, there would probably also be differences between the friction coefficient that is present when the nut is tightened on the washer and the bolt head onto the joint. This would increase the variability still further.
Washers are occasionally used as a means of minimising frictional scatter besides the common reason of reducing the bearing stress on the joint face. The friction condition between the washer and nut face can be reasonably well defined and controlled, more so than the joint surfaces usually can. By controlling the friction, the preload can be more reliably achieved. To do this consistently, a close fit is needed on the inside diameter of the washer. One way in which this can be achieved is by the use of a SEMS unit (in which a washer is held captive on the bolt shank). The same can be achieved by using a KEPS unit (a washer being held captive on a nut).
So in general, when using torque control, tightening the bolt by rotating the bolt head or the nut can matter. It is good practice to specify which part should be tightened so that the bolt preload variation is minimised.
To assist the Engineer in overcoming the problems associated with the use of threaded fasteners and bolted joints, Bolt Science has developed a number of computer programs. These programs are designed to be easy to use so that an engineer without detailed knowledge in this field can solve problems related to this subject.
The Use of Two Nuts to Prevent Self Loosening
Many types of old machinery have two nuts on the bolts. A thin nut is frequently used in these applications. Sometimes the thin nut can be observed below the standard thickness nut and on other installations, it's on top. Although it may seem counter-intuitive, the thin nut should go next to the joint and not be put on last. In other applications, for example on column attachments, two standard thickness nuts are frequently used.
In this article the effectiveness of this locking method is investigated and the tightening procedure that should be used if effective locking is to be achieved.
The use of two plain nuts goes back at least 150 years based upon observation of historic machinery. Tightening one nut down and then simply tightening another nut on top of it 
achieves little locking effect. A specific procedure needs to be followed if locking is to be achieved. When a thin and thick nut are used, it may be thought that the thick nut should go next to the joint since this would take the entire load. However, by placing the thin nut on first, when the thick nut is tightened on top of it, the load on the threads of the thin nut are relieved of their load.
The thin nut should be placed on the bolt first. This nut is typically tightened to between 25% to 50% of the overall tightening torque. The second (thick) nut is then placed on the bolt and the thin nut held to prevent rotation by a spanner whilst the thick nut is tightened to the full torque value. The series of diagrams show the effect that the procedure has on forces present between the nuts and in the bolt.
When the thick nut is tightened onto the thin nut, as the load increases, the load is lifted from the pressure flanks of the thin nut. As tightening continues a point is reached when the bolt thread touches the top flanks of the thin nut. At this point F3 = F2. Continuing to tighten the top nut results in the jamming of the threads leading to F3 > F2. If tightening is continued, the force between the two nuts will continue to increase. If the thick nut is overtightened, there is the risk of thread stripping or the tensile fracture of the bolt between the two nuts.
The reason why the two nut system is effective in resisting self loosening is due to the way the threads are jammed together (hence the term jam nut being frequently used for the thin nut). Since the bolt thread is in contact with the top flank of the small nut and the bottom flank of the top nut, relative thread movement is not possible. For self-loosening to occur, relative movement between the bolt and nut threads must occur. It is this jamming action that is the secret of the two-nut method.
In order to achieve the appropriate bolt preload prior to the threads jamming it is necessary to tighten the smaller nut. The greater the grip length of the joint, the greater is the extension needed to achieve a given preload and hence the higher the initial load that must be sustained by the small nut. Although the axial backlash can be calculated for given tolerance conditions of the nut and bolt threads, there can be a factor of 10 difference between the minimum and maximum values. Such variation makes it difficult to establish the correct preloading of the small nut. As a result, the bottom nut is tightened to a simple percentage (i.e. 25% to 50% of the overall torque value). Two full height nuts can be used if the principles that have been outlined above are followed. Small (jam nuts) are frequently used since there is no need to have a full height nut on the bottom since the threads do not carry the load. An advantage of a thin nut in this application is that a greater amount of axial backlash will be provided for a given tolerance class.
The two videos shown below presents the results of a Junker fastener vibration test performed on the two arrangements that a thick and thin nut on can be arranged. The tests were conducted to investigate the effectiveness of the two-nut method in terms of resistance to self-loosening. A Junker transverse vibration test machine was used with M10 nuts and bolts. The results are illustrated in the graph below. With the small nut on top, both nuts can be observed to rotate together and can subsequently come completely loose. The results are slightly better than is normally observed with a single plain nut. With the small nut next to the joint, some relaxation occurs but not a significant amount of self-loosening . The performance of the two-nut method, when properly applied, provides a superior locking capability when compared to many so-called lock nuts. The proper application of the two-nut method is time intensive and requires a degree of skill and is hence unlikely to make a major comeback on new machinery any time soon.
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