Bolted joints are one of the most common elements in construction and machine design. They consist
of fasteners that capture and join other parts, and are secured with the mating of screw threads.
There are two main types of bolted joint designs: tension joints and shear joints.
In the tension joint, the bolt and clamped components of the joint are designed to transfer
the external tension load through the joint by way of the clamped components through the
design of a proper balance of joint and bolt stiffness. The joint should be designed such
that the clamp load is never overcome by the external tension forces acting to separate
the joint. The second type of bolted joint transfers
the applied load in shear on the bolt shank and relies on the shear strength of the bolt.
Tension loads on such a joint are only incidental. A preload is still applied but is not as critical
as in the case where loads are transmitted through the joint in tension. Other such shear
joints do not employ a preload on the bolt as they allow rotation of the joint about
the bolt, but use other methods of maintaining bolt/joint integrity. This may include clevis
linkages, joints that can move, and joints that rely on a locking mechanism.
Proper joint design and bolt preload provides useful properties:
For cyclic tension loads, the fastener is not subjected to the full amplitude of the
load; as a result, the fastener's fatigue life is increased or—if the material exhibits
an endurance limit its life extends indefinitely. As long as the external tension loads on a
joint do not exceed the clamp load, the fastener is not subjected to motion that would loosen
it, obviating the need for locking mechanisms. For the shear joint, a proper clamping force
on the joint components prevents relative motion of those components and the fretting
wear of those that would result in fatigue cracks.
In both the tension and shear joint design cases, some level of tension preload in the
bolt and resulting compression preload in the clamped components is essential to the
joint integrity. The preload target can be achieved by applying a measured torque to
the bolt, measuring bolt extension, heating to expand the bolt then turning the nut down,
torquing the bolt to the yield point, testing ultrasonically or by a certain number of degrees
of relative rotation of the threaded components. Each method has a range of uncertainties associated
with it, some of which are very substantial.
Theory Typically, a bolt is tensioned by the application
of a torque to either the bolt head or the nut. The preload developed in a bolt is due
to the applied torque and is a function of the bolt diameter, length, the geometry of
the threads and the coefficients of friction that exist in the threads and under the bolt
head or nut. The stiffness of the components clamped by the bolt has no relation to the
preload that is developed by the torque. The relative stiffness of the bolt and the clamped
joint components do, however, determine the fraction of the external tension load that
the bolt will carry and that in turn determines preload needed to prevent joint separation
and by that means to reduce the range of stress the bolt experiences as the tension load is
repeatedly applied. This determines the durability of the bolt when subjected to repeated tension
loads. Maintaining a sufficient joint preload also prevents relative slippage of the joint
components that would produce fretting wear that could result in a fatigue failure of
those parts when subjected to in-plane shearing forces.
The clamp load, also called preload, of a fastener is created when a torque is applied,
and so develops a tensile preload that is generally a substantial percentage of the
fastener's proof strength. A fastener is manufactured to various standards that define, among other
things, its strength and clamp load. Torque charts are available to identify the required
torque for a fastener based on its property class or grade.
When a fastener is torqued a tension preload develops in the bolt and a compressive preload
develops in the parts being fastened. This can be modeled as a spring-like assembly that
has some assumed distribution of compressive strain in the clamped joint components. As
the tension load is applied it relieves the compressive strains induced by the preload,
hence the preload acting on the compressed joint components provides the external tension
load a path other than through the bolt. As long as the forces acting on the fastened
parts do not exceed the preload, the fastener is not subjected to an increase in its tension
load. This however, is a simplified model that is
only valid when the fastened parts are much stiffer than the fastener. In reality, the
fastener carries a small fraction of the external load even if that external load does not exceed
the clamp load. When the fastened parts are less stiff than the fastener, this model breaks
down and the fastener is subjected to a tension load that is the sum of the tension preload
and the external tension load. In some applications, joints are designed
so that the fastener eventually fails before more expensive components. In this case, replacing
an existing fastener with a higher strength fastener can result in equipment damage. Thus,
it is generally good practice to replace old fasteners with new fasteners of the same grade.
Setting the torque Engineered joints require the torque to be
chosen to provide the correct preload. Applying the torque to fasteners is commonly achieved
using a torque wrench. The required torque value for a particular fastener application
may be quoted in the published standard document, defined by the manufacturer or calculated.
A common relationship used to calculate the torque for a desired preload takes into account
the thread geometry and friction in the threads and under the bolt head or nut. The following
assumes standard ISO or National Standard bolts and threads are used:
where is the required torque
is the nut factor is the desired preload
is the bolt diameter The nut factor K accounts for the thread geometry,
friction, pitch. When ISO and Unified National Standard threads are used the nut factor is:
where = the mean thread diameter, close to pitch
diameter. = nominal bolt diameter
=/(pi * dm) Thread Pitch = 1/N where N is the number of
threads per inch or mm = friction coefficient in the threads
= half the thread angle = 30° = friction coefficient under torqued head
or nut When a value of = =0.15 is used and the dimensions
for any size bolt whether course or fine the nut factor is K ≈ 0.20 and the torque/preload
relationship becomes
A study of the effect of torquing two samples, one lubricated and the other unlubricated,
1/2 in.- 20 UNF bolts to 800 lb-in, produced the same mean preload of 7700 lbf. The preloads
for the unlubricated bolt sample had a standard deviation from the mean value of 1100 lbf,
whereas the lubricated sample had a standard deviation of 680 lbf. If the preload value
and torques are used in the above relation to solve for the nut factor it is found to
be K = 0.208, which is very close to the recommended value of 0.20
The preferred bolt preload for structural applications should be at least 75% of the
fastener's proof load for the higher strength fasteners and as high as 90% of the proof
load for permanent fasteners. To achieve the benefits of the preloading, the clamping force
must be higher than the joint separation load. For some joints, multiple fasteners are required
to secure the joint; these are all hand tightened before the final torque is applied to ensure
an even joint seating. The torque value is dependent on the friction
produced in the threads and under the torqued bolt head or nut and the fastened material
or washer if used. This friction can be affected by the application of a lubricant or any plating
applied to the threads, and the fastener's standard defines whether the torque value
is for dry or lubricated threading, as lubrication can reduce the torque value by 15% to 25%;
lubricating a fastener designed to be torqued dry could over-tighten it, which may damage
threading or stretch the fastener beyond its elastic limit, thereby reducing its clamping
ability. Either the bolt head or the nut can be torqued.
If one has a larger bearing area or coefficient of friction it will require more torque to
provide the same target preload. Fasteners should only be torqued if they are fitted
in clearance holes. Torque wrenches do not give a direct measurement
of the preload in the bolt. Much of the torque applied is lost overcoming friction under
the torqued bolt head or nut and in the threads. The remaining 10% of the applied torque does
useful work in stretching the bolt and providing the preload. Initially, as the torque is applied,
it must overcome static friction under the head of the bolt or nut and also in the threads.
Finally, dynamic friction prevails and the torque is distributed in a 5010 manner as
the bolt is tensioned. More accurate methods for determining the
preload rely on defining or measuring the screw extension from the nut. Alternatively,
measurement of the angular rotation of the nut can serve as the basis for defining screw
extension based on the fastener's thread pitch. Measuring the screw extension directly allows
the clamping force to be very accurately calculated. This can be achieved using a dial test indicator,
reading deflection at the fastener tail, using a strain gauge, or ultrasonic length measurement.
Bolt preload can also be controlled by torquing the bolt to the point of yielding. Under some
circumstances, a skilled operator can feel the drop off of the work required to turn
the torque wrench as the material of the bolt begins to yield. At that point the bolt has
a preload determined by the bolt area and the yield strength of the bolt material. This
technique can be more accurately executed by specially built machines. Because this
method only works for very high preloads and requires comparatively expensive tooling,
it is only commonly used for specific applications, primarily in high performance engines.
There is no simple method to measure the tension of a fastener already in place other than
to tighten it and identify at which point the fastener starts extending. This is known
as re-torqueing. An electronic torque wrench can be used on the fastener in question, so
that the torque applied can be constantly measured as it is slowly increased in magnitude.
Recent developments enable tensions to be estimated by using ultrasonic testing. Another
way to ensure correct tension involves the use of crush-washers. These are washers that
have been drilled and filled with orange RTV. When the orange rubber strands appear, the
tension is correct. Large-volume users frequently use computer
controlled nut drivers. With such machines, the computer in effect plots a graph of the
torque exerted. Once the torque reaches a set maximum torque chosen by the designer,
the machine stops. Such machines are often used to fit wheelnuts and normally tighten
all the wheel nuts simultaneously. Thread engagement
Thread engagement is the length or number of threads that are engaged between the screw
and the female threads. Screws are designed so that the bolt shank fails before the threads,
but for this to hold true, a minimum thread engagement must be used. The following equation
defines this minimum thread engagement:
Where Le is the thread engagement length, At is the tensile stress area, D is the major
diameter of the screw, and p is the pitch. This equation only holds true if the screw
and female thread materials are the same. If they are not the same, then the following
equations can be used to determine the additional thread length that is required:
Where Le2 is the new required thread engagement. While these formulas give absolute minimum
thread engagement, many industries specify that bolted connections be at least fully
engaged. For instance, the FAA has determined that in general cases, at least one thread
must be protruding from any bolted connection. [1]
Failure modes The most common mode of failure is overloading:
Operating forces of the application produce loads that exceed the clamp load, causing
the joint to loosen over time or fail catastrophically. Overtorquing might cause failure by damaging
the threads and deforming the fastener, though this can happen over a very long time. Undertorquing
can cause failures by allowing a joint to come loose, and it may also allow the joint
to flex and thus fail under fatigue. Brinelling may occur with poor quality washers,
leading to a loss of clamp load and subsequent failure of the joint.
Other modes of failure include corrosion, embedment, and exceeding the shear stress
limit. Bolted joints may be used intentionally as
sacrificial parts, which are intended to fail before other parts, as in a shear pin.
Locking mechanisms
Locking mechanisms keep bolted joints from coming loose. They are required when vibration
or joint movement will cause loss of clamp load and joint failure, and in equipment where
the security of bolted joints is essential. Two nuts, tightened on each other. In this
application a thinner nut should be placed adjacent to the joint, and a thicker nut tightened
onto it. The thicker nut applies more force to the joint, first relieving the force on
the threads of the thinner nut and then applying a force in the opposite direction. In this
way the thicker nut presses tightly on the side of the threads away from the joint, while
the thinner nut presses on the side of the threads nearest the joint, tightly locking
the two nuts against the threads in both directions. Bolt banging
Bolt banging occurs in buildings when bolted joints slip into bearing under load, thus
causing a loud and potentially frightening noise resembling a rifle shot that is not,
however, of structural significance and does not pose any threat to occupants.
International standards SA-193/SA-193M: "Specification for alloy-steel
and stainless steel bolting materials for high-temperature service"
SA-194/SA-194M: "Specification for carbon and alloy steel nuts for bolts for high-temperature
service" SA-320/SA-320M: "Specification for alloy steel
bolting materials for low-temperature service" EN 1515: "Flanges and their joints - Bolting"
EN 1515-1: "Flanges and their joints - Bolting - Part 1: Selection of bolting"
EN 1515-2: "Flanges and their joints — Bolting — Part 2: Classification of bolt materials
for steel flanges, PN designated" EN 1515-2: "Flanges and their joints — Bolting
— Part 3: Classification of bolt materials for steel flanges, class designated"
ISO 4017: "Hexagon head screws - Product grades A and B"
ISO 4032: "Hexagon nuts, style 1 - Product grades A and B"
ISO 4033: "Hexagon nuts, style 2 - Product grades A and B"
See also Bearing surface
Bolt manufacturing process Castellated nut/capscrew
Quenching and tempering Rivet
Locknut Polymer insert nut
Oval lock nut
Lock washer Lock wire
Mechanical joint Thread adhesive
Residual stress References
Notes
Bibliography Collins, Jack A.; Staab, George H.; Busby,
Henry R., Mechanical Design of Machine Elements and Machines, Wiley, ISBN 0-471-03307-3.
Oberg, Erik; Jones, Franklin D.; McCauley, Christopher J.; Heald, Ricardo M., Machinery's
Handbook, Industrial Press, ISBN 978-0-8311-2700-8. External links
Bolt Length Calculation Method Bolts - AISC | Home
AISC THE BANGING BOLT SYNDROME AISC BANGING BOLTS— ANOTHER PERSPECTIVE
Bolt Science - The Jost Effect Threaded Fasteners - Tightening to Proper
Tension, US Department of Defense document MIL-HDBK-60, 2.6MB pdf.
NASA Reference Publication 1228 Fastener Design Manual
Mechanics of screws FAA Advisory Circular 43.13-1B, Paragraph
7-37 "Grip Length" Bolted Joint Design, Fastenal Engineering
& Design Support