FAQ
You can find more information on our web, so please take a look.
Questions we are Frequently Asked
Some of the frequently asked questions we get asked are presented below:
What
are the marks shown on the head of a bolt?
When
tightening stainless steel bolts - they tend to seize - what's
happening?
I
can't find the shear strength of a fastener in the specification,
can you help?
What
is the best way to check the torque value on a bolt?
What
are the benefits of fine threaded fasteners over coarse threaded
fasteners?
What
methods are available for calculating the appropriate tightening
torque for a bolt.
Does
it matter whether you tighten the bolt head or the nut?
How
do you select a fastener size for a particular application?
Does
using an extension on a torque wrench change the abliity to
achieve the desired torque value?
Is
it okay to use a mild steel nut with a high tensile bolt?
Should
I always use a washer under the bolt head and nut face?
What
is the torque to yield tightening method?
How
do metric strength grades correspond to the inch strength
grades?
What
is the difference between a bolt and a screw?
Are
the use of a thin nut and a thick nut effective in preventing
loosening?
Is
there some standard that states how much the thread should
protrude past the nut?
Some
bolts are deliberately tightened past their yield point. Why
don't they further yield when an external load is subsequently
applied to the joint and come loose?
What are the marks shown on the head of a bolt?
Usually fastener standards specify two types of marks to be on the head of a bolt. The manufacturer's mark is a symbol identifying the manufacturer (or importer). This is the organisation that accepts the responsibility that the fastener meets specified requirements. The grade mark is a standardised mark that identifies the material properties that the fastener meets. For example 307A on a bolt head indicates that the fastener properties conform to the ASTM A307 Grade A standard. The bolt head shown at the side indicates that it is of property class 8.8 and ML is the manufacturer's mark.
Both marks are usually located on the top of the bolt head,
most standards indicating that the marks can be raised or
depressed. Raised marks are usually preferred by manufacturers
because these can only be added during the forging process
whereas depressed marks can subsequently added (possibly with
illegitimate marks).
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We have a problem when tightening stainless steel bolts - they tend to seize - whats happening?
Stainless steel can unpredictably sustain galling (cold welding). Stainless steel self-generates an oxide surface film for corrosion protection. During fastener tightening, as pressure builds between the contacting and sliding, thread surfaces, protective oxides are broken, possibly wiped off, and interface metal high points shear or lock together. This cumulative clogging-shearing-locking action causes increasing adhesion. In the extreme, galling leads to seizing - the actual freezing together of the threads. If tightening is continued, the fastener can be twisted off or its threads ripped out.
If galling is occurring than because of high friction the
torque will not be converted into bolt preload. This may be
the cause of the problems that you are experiencing. The change
may be due to the surface roughness changing on the threads
or other similar minor change. To overcome the problem - suggestions
are:
1. Slowing down the installation RPM speed may possibly solve
or reduce the frequency of the problem. As the installation
RPM increases, the heat generated during tightening increases.
As the heat increases, so does the tendency for the occurrence
of thread galling.
2. Lubricating the internal and/or external threads frequently
can eliminate thread galling. The lubricants usually contain
substantial amounts of molybdenum disulfide (moly). Some extreme
pressure waxes can also be effective. Be careful however,
if you use the stainless steel fasteners in food related applications
some lubricants may be unacceptable. Lubricants can be applied
at the point of assembly or pre-applied as a batch process
similar to plating. Several chemical companies, such as Moly-Kote,
offer anti-galling lubricants.
3. Different combinations of nut and bolt materials can assist
in reducing or even eliminating galling. Some organisations
specify a different material, such as aluminium bronze nuts.
However this can introduce a corrosion problem since aluminium
bronze is anodic to stainless steel.
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I can't find the shear strength of a fastener in the specification, can you help?
Bolted shear joints can be designed as friction grip or direct shear. With friction grip joints you must ensure that the friction force developed by the bolts is sufficient to prevent slip between the plates comprising the joint. Friction grip joints are preferred if the load is dynamic since it prevents fretting.
With direct shear joints the shank of the bolts sustain the shear force directly giving rise to a shear stress in the bolt. The shear strength of a steel fastener is about 0.6 times the tensile strength. This ratio is largely independent of the tensile strength. The shear plane should go through the unthreaded shank of a bolt if not than the root area of the thread must be used in the calculation.
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What is the best way to check the torque value on a bolt?
There are three basic methods for the checking of torques applied to bolts after their installation; namely, taking the reading on a torque gauge when:
1. The socket begins to move away from the tightened position in the tightening direction. This method is frequently referred to as the "crack-on" method.
2. The socket begins to move away from the tightened position in the un- tightening direction. This method is frequently referred to as the "crack-off" method.
3. The fastener is re-tightened up to a marked position. With the "marked fastener" method the socket approaches a marked position in the tightening direction. Clear marks are first scribed on the socket and onto the joint surface which will remain stationary when the nut is rotated. (Avoid scribing on washers since these can turn with the nut.) The nut is backed off by about 30 degrees, followed by re-tightening so that the scribed lines coincide.
For methods 1. and 2. the breakloose torque is normally slightly higher than the installation torque since static friction is usually greater than dynamic friction. In my opinion, the most accurate method is method 3 - however what this will not address is the permanent deformation caused by gasket creep. An alternative is to measure the bolt elongation (if the fastener is not tapped into the gearbox). This can be achieved by machining the head of the bolt and the end of the bolt so that it can be accurately measured using a micrometer. Checking the change in length will determine if you are losing preload.
The torque in all three methods should be applied in a slow and deliberate manner in order that dynamic effects on the gauge reading are minimised. It must always be ensured that the non- rotating member, usually the bolt, is held secure when checking torques. The torque reading should be checked as soon after the tightening operation as possible and before any subsequent process such as painting, heating etc. The torque readings are dependent upon the coefficients of friction present under the nut face and in the threads. If the fasteners are left to long, or subjected to different environmental conditions before checking, friction and consequently the torque values, can vary. Variation can also be caused by embedding (plastic deformation) of the threads and nut face/joint surface which does occur. This embedding results in bolt tension reduction and affects the tightening torque. The torque values can vary by as much as 20% if the bolts are left standing for two days.
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What are the benefits of fine threaded fasteners over coarse threaded fasteners?
The potential benefits of fine threads are:
1. Size for size a fine thread is stronger than a coarse thread . This is both in tension (because of the larger stress area) and shear (because of their larger minor diameter).
2. Fine threads have also less tendency to loosen since the thread incline is smaller and hence so is the off torque.
3. Because of the smaller pitch they allow finer adjustments in applications that need such a feature.
4. Fine threads can be more easily tapped into hard materials and thin walled tubes.
5. Fine threads require less torque to develop equivalent bolt preloads.
On the negative side:
1. Fine threads are more susceptible to galling than coarse threads.
2. They need longer thread engagements and are more prone to damage and thread fouling.
3. They are also less suitable for high speed assembly since they are more likely to seize when being tightened.
Normally a coarse thread is specified unless there is an over-riding reason to specify a fine thread, certainly for metric fasteners, fine threads are more difficult to obtain.
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What methods are available for calculating the appropriate tightening torque for a bolt?
A high bolt preload ensures that the joint is resistant to vibration loosening and to fatigue. In most applications, the higher the preload - the better (assuming that the surface pressure under the nut face is not exceeded that is).
The preload is related to the applied torque by friction that is present under the nut face and in the threads. The torque value depends primarily on the values of the underhead and thread friction values and so a single figure cannot be quoted for a given thread size.
The stress that is often quoted is often taken as the direct stress in the bolt as a result of the preload. It is normally calculated as preload divided by the stress area of the thread. Typical values vary between 50% to 80% of the yield strength of the bolt material, in many applications a figure of 75% of yield is used. Our TORKSense program uses this approach and further details on this is presented in the help file that comes with the demo program that is available for download from our web site. (This program also provides large databases on thread, bolting materials and nut factors.)
It is important to note that it does not take into account the torsional stress as a result of the tightening torque. High friction values can push the actual combined stress over yield if high percentages are used. (The tensile stress from the preload coupled with a high torsional shear stress from the torque due to thread frictional drag results in a high combined stress.) The percentage yield approach works well in most practical circumstances but if you are using percentage of yield values over 75% then you could be exceeding yield if high friction values are being used.
One way to over come this limitation is to use the percentage of yield based upon the combined effects of the direct stress (from the bolt preload) and the torsional stress (from the applied torque). Using this approach to specify torque values is more logically consistent and can reduce the risk of the yield strength of the bolt being exceeded - especially under high thread friction conditions. A figure of 90% of yield is typically used here when the combined stress (usually calculated as the Von-Mises stress) from the direct and torsional stresses is calculated. Our Torque and BOLTCALC programs uses this approach and a copy of the demo program can be downloaded from our web site. The help file provided with the demo program does provide additional information on this topic.
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Does it matter whether you tighten the bolt head or the nut?
Normally it will not matter whether the bolt head or the nut is torqued. This assumes that the bolt head and nut face are of the same diameter and the the contact surfaces are the same (giving the same coefficient of friction). If they are not then it does matter.
Say the nut was flanged and the bolt head was not. If the tightening torque was determined assuming that the nut was to be tightened then if the bolt head was subsequently tightened instead then the bolt could be overloaded. Typically 50% of the torque is used to overcome friction under the tightening surface. Hence a smaller friction radius will result in more torque going into the thread of the bolt and hence being over tightened.
If the reverse was true - the torque determined assuming that the bolt head was to be tightened then if the nut was subsequently tightened - the bolt would be under tightened.
There is also an effect due to nut dilation that can, on occasion, be important. Nut dilation is the effect of the external threads being pushed out due to the wedge action of the threads. This reduces the thread stripping area and is more prone to happen when the nut is tightened since the tightening action facilitates the effect. Hence if thread stripping is a potential problem, and for normal standard nuts and bolts it is not, then tightening the bolt can be beneficial.
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How do you select a fastener size for a particular application?
When selecting a suitable fastener for a particular application there are several factors that must be taken into account. Principally these are:
1. How many and what size/strength do the fasteners need to be? Other than rely upon past experience of a similar application an analysis must be completed to determine the size/number/strength requirements. A program like BOLTCALC can assist you with resolving this issue.
2. The bolt material to resist the environmental conditions prevailing. This could mean using a standard steel fastener with surface protection or may mean using a material more naturally corrosion resistant such as stainless steel.
The general underlying principle is to minimise the cost of the fastener whilst meeting the specification/life requirements of the application. Each situation must be considered on its merit and obviously some detailed work is necessary to arrive at a detailed recommendation.
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Does using an extension on a torque wrench change the abliity to achieve the desired torque value?
If you use an extension spanner on the end of a torque wrench, the torque applied to the nut is greater than that shown on the torque wrench dial.
If the torque wrench has a length L, and the extension spanner a length E (overall length of L+E) than:
TRUE TORQUE= DIAL READING X (L+E)/L
i.e the torque will be increased.
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Is it okay to use a mild steel nut with a high tensile bolt?
Nut thickness standards have been drawn up on the basis that the bolt will always sustain tensile fracture before the nut will strip. 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. Hence, the potential of thread stripping of both the internal and external threads must be avoided if a reliable design is to be achieved. When specifying nuts and bolts it must always be ensured that the appropriate grade of nut is matched to the bolt grade.
The standard strength grade (or Property Class as it is known in the standards) for many industries is 8.8. On the head of the bolt, 8.8 should be marked together with a mark to indicate the manufacturer. The Property Class of the nut matched to a 8.8 bolt is a grade 8. The nut should be marked with a 8, a manufacturer's identification symbol shall be at the manufacturer's discretion.
Higher tensile bolts such as property class 10.9 and 12.9 have matching nuts 10 and 12 respectively. In general, nuts of a higher property class can replace nuts of lower property class (because as explained above, the 'weakest link' is required to be the tensile fracture of the bolt).
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Should I always use a washer under the bolt head and nut face?
Our opinion is that plain washers are best avoided if possible and certainly, a plain washer should not be used with a 'lock' washer. It would partly negate the effect of the locking action and secondly could lead to other problems (see below). Many 'lock' washers have been shown to be ineffective in resisting loosening.
The main purpose of a washer is to distribute the load under the bolt head and nut face. Instead of using washers however the trend as been to the use of flanged fasteners. If you compute the bearing stress under the nut face it often exceeds the bearing strength of the joint material and can lead to creep and bolt preload loss. Traditionally a plain washer (that should be hardened) is used in this application. However they can move during the tightening process (see below) causing problems.
Research indicates that the reason why fasteners come loose is usually caused by transverse loadings causing slippage of the joint. The fastener self loosens by this method. When using impact tightening tools there is a large variability in the preload achieved by the fastener. The tightening factor is between 2.5 and 4 for this method. (The tightening factor is the ratio of max preload to min. preload.) Software such as our BOLTCALC program allow for this by basing the design on the lowest anticipated preload that will be achieved in the assembly. Because of changes in the thread condition itself - different operators etc. it could be that lower values of preload are being achieved even though the assemblies may appear to be identical.
One problem that can occur with washers is that they can move when being tightened so that the washer can rotate with the nut or bolt head rather than remaining fixed. This can affect the torque tension relationship.
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What is the torque to yield tightening method?
Torque to yield is the method of tightening a fastener so that a high preload is achieved by tightening up the yield point of the fastener material. To do this consistently requires special equipment that monitors the tightening process. Basically, as the tightening is being completed the equipment monitors the torque verses angle of rotation of the fastener. When it deviates from a specified gradient by a certain amount the tool stops the tightening process. The deviation from a specified gradient indicates that the fastener material as yielded.
The torque to yield method is sometimes called yield controlled tightening or joint controlled tightening.
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How do metric strength grades correspond to the inch strength grades?
Some details on conversion guidance between metric and inch based strength grades is given in section 3.4 of the standard SAE J1199 (Mechanical and Material Requirements for Metric Externally Threaded Steel Fasteners).
Metric fastener strength is denoted by a property class which is equivalent to a strength grade. Briefly:
Class 4.6 is approximately equivalent to SAE J429 Grade 1 and ASTM A307 Grade A
Class 5.8 is approximately equivalent to SAE J429 Grade 2
Class 8.8 is approximately equivalent to SAE J429 Grade 5 and ASTM A449
Class 9.8 is approximately 9% stronger than equivalent to SAE J429 Grade 5 and ASTM A449
Class 10.9 is approximately equivalent to SAE J429 Grade 8 and ASTM A354 Grade BD
For information there is no direct inch equivalent to the metric 12.9 property class.
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What is the difference between a bolt and a screw?
Historically the difference between a bolt and a screw was that the screw was threaded to the head whereas the bolt had a plain shank. However I would say that now this could cause you a problem if you made this assumption when specifying a fastener. The definition used by the Industrial Fastener Institute (IFI) is that screws are used with tapped holes and bolts are used with nuts.
TRM supply professional and honest service.
Obviously a standard 'bolt' can be used in a tapped hole or with a nut. The IFI maintain that since this type of fastener is normally used with a nut then it is a bolt. Certain short length bolts are threaded to the head - they are still bolts if the main usage is with nuts. Screws are fastener products such as wood screws, lag screws and the various types of tapping screws. The IFI terminology and definition has been adopted by ASME and ANSI.
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Are the use of a thin nut and a thick nut effective in preventing loosening?
I had been of the opinion that when two nuts were being used to lock a thread, the thicker of the two nuts should go next to the joint. I had this as one of the 'tips for the day' on some software and a couple of years ago was taken to task that this was wrong. The thin nut he said should go next to the joint.
My reasoning was that nut heights had been
decided by establishing the least height that would ensure
that the bolt would break before the threads started to shear.
So if you wanted to get the maximum preload into the fastener
then the thick nut should go first so that thread stripping
was prevented. If you put the thin nut first, the preload
would be limited by the thread stripping (whose failure may
not be obvious at the time of the nuts were tightened). Putting
the thin nut on top of the thick nut, I thought, would assist
in preventing the thick nut self-loosening. I had also seen
that
using two nuts was a popular method on old machinery - and
the ones that I had seen all had the thin nut on top of the
thick nut.
The correct procedure, I was told, was
to put the thin nut on first, tighten it to 30% or so of the
full torque and then tighten the thick nut on top of it to
the full torque value. You have to take care that the thin
nut does not rotate when you are tightening the thick nut.
The tightening of the thick nut would impose a preload on
the joint equivalent to that which would be obtained from
100 - 30 = 70% of the tightening torque (approximately anyway).
The idea is that the bolt threads engaging on the thin nut
disengage so that the thick nut takes the preload by taking
up the backlash on the threads of the thin nut. The thin nut
being jammed (hence the alternative name - jam nut) against
the thick
nut. This helps to prevent self-loosening and improves the
fastener's fatigue performance by modifying the load distribution
within the threads. Doing it the other way, thin nut on top
of the thick nut, does not jam the parts together sufficiently.
Two years on and I am still unconvinced. I am still asked the two nut question but I always tend to recommend other more modern ways of locking the threads. I think that the reasons that I am not easy with the method is that it is too reliant upon the skill of the person tightening the joint. There is also the amount of backlash in the threads (you could strip the threads of the small nut if it was a tight fit) and the preload will be down on what it could be as well.
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Is there some standard that states how much the thread should protrude past the nut?
There are some building codes that stipulates that there must be at least one thread protruding through the nut. However it is common practice to specify that at least one thread pitch must protrude across a range of industries. Typically the first few pitches of the thread can be only partially formed because of a chamfer etc.
Nut thickness standards have been drawn up on the basis that the bolt will always sustain tensile fracture before the nut will strip. 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. Hence, the potential of thread stripping of both the internal and external threads must be avoided if a reliable design is to be achieved. When specifying nuts and bolts it must always be ensured that the appropriate grade of nut is matched to the bolt grade.
In cases of when a threaded fastener is tapped into a plate or a block it is usually the case that the fastener and block materials will be of different strengths. If the criteria is adopted that the bolt must sustain tensile fracture before the female thread strips, the length of thread engagement required can be excessive and can become unrealistic for low strength plate/block materials. Tolerances and pitch errors between the threads can make the engagement of long threads problematical.
In summary the full height of the nut is to be used if you are to avoid thread stripping. Have a look at information on the website on the BOLTCALC program and thread stripping - there is a tutorial/presentation available from the website.
In terms of maximum protrusion I have not come across any guidelines on this point other then minimise to avoid wasting material.
See shortbolting.htm for more information on this topic.
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Some bolts are deliberately tightened past their yield point. Why don't they further yield when an external load is subsequently applied to the joint and come loose?
When a bolt is tightened into its plastic region, yielding is the result of the combined effects of both tensile/axial stress and the torsional stress exceeding the yield point of bolt material. The tensile/axial stress is the result of the extension/stretching of the bolt and the torsional stress as a result of the thread friction and stretch torque acting on the thread. When the joint sustains an external loading, there are two effects that allow the bolt to be axially loaded without sustaining further plastic deformation:
1. A significant proportion of the torsion in the bolt, typically
50% or so, disappears as soon as the tightening operation
is concluded. There is a change in the torque reaction within
the fastener. For example, if the nut is tightened, there
will be a torsion acting down the shank being driven from
the socket and reacted at the bolt head. When the socket is
removed, the torsion is then reacted between the nut face
and the bolt head with it being reduced by 50% or so. The
remaining torsion is thought normally to disappear as a result
of embedding/relaxation losses.
2. A new yield point forms at the point
on the strain curve that the bolt had been tightened to. This
effect is referred to as the Bauschinger effect (more details
on this effect is available at https://en.wikipedia.org/wiki/Bauschinger_effect).
The net result of these two effects discussed
above is that even with the bolt tightened plastically it
will perform elastically when external loads are applied to
the joint. Obviously there are limits to the magnitude of
the load that can be applied before yielding occurs. In many
applications, joint separation occurs before the bolt yields.
One important exception is for joints consisting of different
materials subjected to a significant temperature change. One
such application is a steel bolt in an aluminium joint. In
such joints the bolt can further yield as a result of differential
thermal expansion subsequent to it being tightened. (The coefficient
of thermal expansion of aluminium is broadly twice that of
steel and so the joint thickness increases with rising temperature
at a greater rate than the steel expands.) Effects such as
a reduction in the yield strength of the material at an elevated
temperature can also play a part. In some of these cases,
for example cylinder head bolts, yielding can occur when the
engine first starts and heats the block but the yielding is
limited and is stabilised in subsequent heat-cooling cycles.
.
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Question On Slip/Bolting Joint
thread404-421821 Forum Search FAQs Links MVPsForum
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MVPs
(Industrial)
(OP)
2 Mar 17 10:18Just a quick question, and grateful for any guidance.I have an simple application that has 3 Bolts all have the same diameter, d, and has a slight clearance with the hole, D. They are responsible for clamping 2 plates together, or it could be 3 dependent on design.They are tensioned, and a clamping force is exterted on the 2/3 plate surfaces, F.A axial shear force, P is then applied to the joint.The Plates are made from Steel.The Shear Force Per Bolt = P / 3 (No. Bolts)The Total Clamping Force on joint is = F x 3If the joint was to move and slip and enter a bearing situation, then the Axial Shear Force has to be greater than the Total Clamping Force.However, as there will be a co-efficient of friction between the plates i.e. 0.33 (Steel/Steel), then I presume that the total force to stop slippage would be Shear Load, P * Coefficient of FrictionTherefore if the Force to move the Joint that is under the Clamping Load must, then P x Co Eff Friction >= F.Am I correct in my summation?
(Mechanical)
2 Mar 17 15:26Hello,
Wouldn't your total clamping force be F, not 3 x F?
I dont think adding more bolts increases the clamping force if they are all torqued the same.
(Mechanical)
2 Mar 17 16:36daparojo,Your assumption is correct. Your sliding force P=3Fμ. This is Newtonian friction. I am assuming your bolts are close together such that there are no weird moments developing.Assume your coefficient of friction is not very accurate. Assume your bolt tension due to torque is not very accurate. Assign factors of safety accordingly.
--
JHG
(Chemical)
2 Mar 17 19:42I do not know what an axial shear force is and your drawing does not show it correctly. Your drawing shows some assembly that will be accelerating off to the left. In Statics a free-body diagram should be used, which will show all of the forces and they will add up to zero.
The bolts should not carry any shear load unless they are in contact with both plates at the interface and are being sheared, as in a a pair of scissors. The bolts are in tension and the shear loads are transferred through friction between the plates.
There is no way to calculate the actual shear on these bolts because it depends entirely on the fit and tolerances of the whole assembly. For calculation purposes it is assumed the shear is zero, and that the shear is transferred through friction.
(Mechanical)
2 Mar 17 21:27HiBecause one cannot guarantee contact between the bolts and the hole edge the shear force is resisted by the friction between the clamped plates. So force to overcome friction is the total clamping force provided by the bolts multiplied by the friction coefficient.
“Do not worry about your problems with mathematics, I assure you mine are far greater.” Albert Einstein
(Industrial)
(OP)
2 Mar 17 22:30Cheers guys.
Desertfox, if the Shear Force was 500 kN and the total Clamping Load from the Bolts was 300 kN i.e 100 kN per Bolt.
Take CoF 0.5
I am confused, please bear with me. The way I saw it...
The shear force on the plates with friction from the plates would be 500 kN × 0.5 = 250 kN
As the clamping force is greater than the force trying to move it, the joint wouldnt move?
But as you put it, 300 kN x 0.5 = 150kN across the plate which is less than the original clamping force? How can the joint lose clamping force ?
The way I tried to visualize what was happening. If I held a piece of plate between my fingers, and exerted little force on the surface, then it would take little force to overcome the friction to cause it to slip through my fingers. If I then put more pressure on the plates with my fingers, it would take more force to move the plates. Thats why I am finding it hard to understand that, say 300 kN clamping load x 0.5 CoF is less clamping pressure on the plates. Taking 0.5 as roughish surface CoF If the CoF was 1 as Glass/Glass it would take more force to move the plates.
Thanks for you understandmg
(Mechanical)
2 Mar 17 23:40(shear force) = (faying surface friction force) <= ( (clamp force) * (coefficient of friction) )
daparojo,I'm basically restating what others have already said, but the following must be true to prevent slip:You had it backwards in your original post. Any coefficient of friction less than one will mean the shear force required to cause slip will be less than the clamp force. Look up "slip-critical joint" or "slip-critical connection" for more information on the specific topic, but you may want to read more about friction in general first.
pylfrm
(Mechanical)
3 Mar 17 05:36HiThe joint hasn't lost clamping force but if you have a total bolt clamping force of 300KN then the shear force acting at right angles to overcome the the bolt clamping is resisted only by friction, therefore in accordance with the sliding friction lawmu = F/RF is the shear force in your case.R = the normal reaction ( this is the bolt clamping force)Imagine having to push a large heavy crate say 100kg, to move it one only as to overcome sliding friction, so if mu is 0.25 then exerting 25kg will slide the crate but to lift it you would need at least 100kg force
“Do not worry about your problems with mathematics, I assure you mine are far greater.” Albert Einstein
(Industrial)
(OP)
3 Mar 17 10:12Cheers guys, and the help is much appreciated.
I can fulled understand that say a 100Kg load with a mu=0.25 would take 25 Kg of force to move it, but if mu=1 then it would take 100Kg of force to move something on a more rougher surface - makes sense.
Its been a confusion that an application from 30 + years ago that I am looking and have found that the clamping force on the 3 off bolts under friction is alot less than the shear force that can be acting on it - therefore it will move. However, this has been deemed accpetable years ago!
(Mechanical)
3 Mar 17 15:00Hi daparojoyou're very welcome.
“Do not worry about your problems with mathematics, I assure you mine are far greater.” Albert Einstein
(Nuclear)
3 Mar 17 16:27Over that long a service history, are you sure the original assumed torques are correct? Has the joint ever actually been twisted that much?
Or has it twisted under use EVERY time and the twist torque forces are actually being resisted by the 3x bolts rotating in their bolt holes to act as shear pins?
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