okay uh
some of the other methods of direct
reading of fastener tension
load cells
and of course this
pli
preliminar or preload indicating washer
that i described in the washer section
is a mechanical load cell assembly and
those can be used
but of course in a lot of assemblies you
don't have room enough to put all that
stuff in
because the amount of deformation in
them has been correlated to specific
tension load in the bolt
now the skidmore will helm bolt tension
measuring machine
shown on the next page has a load cell
in it to give a direct bolt tension
reading for an applied torque now it's a
bench top type
uh setup which can be
used at a construction site you just
clamp it on to a beam and check some
bolts and then you
determine the torque that you want to
put in that particular batch of bolts
and use that for an installation
this company is uh here in uh cleveland
uh
the uh one of the owners of it attended
uh last uh fastener conference we had
here
see it's a little job it just has the
the clamps over here that you can clamp
it on and it gives a direct reading of
the number of pounds that you've put
into the bolt so it
it is a way
of determining
uh torque
now here is
here's another one that you don't think
of it as being a uh
direct reading but it is and i use this
one i'll cover it later in the rivets
section this is a high lock rivet
and uh
or a lock bolt it's actually a lock bolt
pardon me rather than a rivet
because it the shank does not expand on
it
but it has a
it's a blind type fastener
for installation and you have a hex key
that holds it in place
while you crank the nut on and when you
get the nut to the proper torque limit
it is notched here
so that it snaps off breaks off so that
is a self-limiting type
that they have determined the
right diameter for them
so it will break off at the torque that
you want
now here's one that i mentioned earlier
the the dti bolt
and the guy who has that company was at
our
one of our building technology council
meetings
this is a
colored coated type bolt it has a
little
gauge pin that's threaded inside it
and then it has a
an optical absorptance cell near the
surface
and as the cell changes thickness it
changes color
so
as you
elongate the bolt it pulls this gauge
away from the cell
and gives you a color-coded load
indication so you turn the thing until
it shows red or whatever and you're all
right but of course
the problem with a bolt like this you
can imagine how expensive it is compared
to a hardware store bolt
and it's got to be big enough that you
can drill the center of it to put that
stuff in so the minimum size for it is a
half inch diameter
now we move on to design criteria
and
this first line is one of my pet peeves
i think that we don't spend enough time
looking at something before we design it
we should think about it look at it
first now of course working with the
research people around here you usually
do
designs by an iterative process because
when they come to you about something
usually they don't know and i'm not
saying this as a
degrading remark they don't know exactly
what they want so they tell you what
they think they want and then you take
it from there and usually what happens
is you by iterative process
you come up with the actual requirements
and that's something that you should do
on any design you should sit down first
and look at it
decide what you really need
and then look at the accepted design
practices from both the layout and
analytical standpoints
to see what you should do
now here's here's one of the questions
that comes up sometimes is diameter
versus length on fasteners
and uh
we're always faced with decisions on do
you use off-the-shelf stuff
or do you custom design it
well
a good way to look at it is to check and
see what's available first see if you
can
build your design around that without
having to buy special
components
and
so
one of the things that
you look at is the length to diameter
ratio
of
because if you want to use a 10 inch
fastener that's a quarter inch in
diameter
and you don't want to make it out of
threaded rod
you're in trouble because you're going
to have to get one that's custom made
usually the l over d ratio is up to
about 12
and uh it's limited somewhat by the
capacity of the automatic screw forming
machines because you can't put a long
skinny fastener through there and do it
on an automated basis
so we have a table here that lists
common
fasteners
availability now these are industrial
fasteners
not aerospace
and
the
one asterix
represents the stock sizes of maximum
demand so you see if you're in this this
area if you need a
3 8 diameter
with an inch or inch and a quarter
length you got it
two asterisks represents the ones less
frequently used so if you want say a
quarter inch and an inch and three
quarter
length you might have a little trouble
all the rest of them are considered
specials
so if you want a quarter inch
by six inch look look how deep of
trouble you're in down here because you
can't get it unless you pay special for
it
and i know we had some
fasteners on a job that we did here one
time that were quarter inch stainless
steel that had to be i believe
six inches or five and a half inches or
six inches long
we had to pay special for them and took
a long time to get them now when i say
these are industrial type fasteners on
the aerospace fasteners the links there
are graduated in sixteenths
uh the different specified dash number
that gives you the
grip length of the fastener in
sixteenths but just because somebody
shows it in their catalog doesn't mean
that they have it either so if you need
something that's an oddball type uh
length to diameter you still are going
to have to to pay special for it now
here's a little handy dandy thing that
uh once again i
got it from one of these guys at martin
when i worked there
and i've never seen this anywhere either
is a way of calculating the number
fastener diameter
i put it in my fastener manual so a lot
of you guys have already seen it but
in order to calculate it this is for
inches of course you take 60 thousandths
plus 13 thousandths times n where n is
the number of the fastener
and that'll give you the total od of it
and i the example i give here is a
number eight fastener you take sixty
thousandths plus thirteen thousands
times eight gives you a 164. so if
somebody says i got a number six
or a number four you can calculate the
decimal diameter of it directly that way
just by keeping that in mind of course
the number 10 is easy
because 130 plus 60 is 190.
and
for those of you who are
haven't seen them before they even have
number 12 fasteners
which is i believe works out to be 0.216
or something like that the automotive
industry uses them some and i ran into
some of them one time and i couldn't
figure out what they were way back years
years ago until i figured out it was a
number 12
but they're they're not a normal one
now clearance holes for fasteners
for shear applications
the clearance should be minimized
and uh ideally the hole should be max
drilled
and the material thickness and fastener
strength should be sized to make the
fasteners critical in bearing rather
than shear that means that if you pull
and shear on the joint
that the
fastener is stronger than the material
it's in so it will elongate the hole so
it can load up the other fasteners
and uh in tension applications you don't
have to worry about that if it's if you
can assure yourself that you have enough
tension in it enough friction that the
joint won't move
then you can have a looser fit on it
then then your main concern is prevent
the fastener head or the nut from
pulling through the hole or something
like that or embedding in it now on the
next page
fred yaris's group was kind enough to
draw me up this little thing because i
couldn't find one usually you try to
steal stuff from someplace else for
these to save yourself work but
uh
we couldn't get by with it so we had to
make one
and this is a little
drawing of a joint to illustrate
the clearance hole gaps on fasteners and
where it gets you in trouble
now this happens a lot
i'll go to this one over here to maybe a
little clearer this happens a lot where
you have two pieces
that
one place makes one piece and somebody
else makes the other one then you bring
them back and you try to put them
together
and this is what you get now these are
the different gaps see
see here we have no gap on this
but look what we got up here and look
what we got here
so if you pull on that
the only way
that these other fasteners can load up
like for instance here
since that one is up against the the
wall right now
in order for this one to load up
the hole has to elongate on here for
that one to load up so
so this is why that in a
real shear applications critical design
you should match drill
and this is what the aerospace companies
do they'll take the pieces
they'll have a pilot hole in one
which is a smaller diameter hole than
the the hole that needs to be in it at
the end
they will clamp them together then they
will use that pilot hole
to go in with the proper size drill and
drill the hole all the way through both
pieces so that it matches perfectly
drilled with the same drill
and then you put it together and you
don't have this problem
now here here's another one that
we can run into trouble on is mixing of
the thread and material types and this
happens in designs sometimes
because
you can have say 300 series stainless
steel fasteners
and you can have a286 stainless steel
fasteners and you look at them
they look alike
the one has a strength of usually of 160
and the other one has a strength of 70.
so you can get in trouble with it so
if the different sizes
have fine or coarse threads on the same
diameters or i mean if you have them
with the same diameters with the
finer coarse threads or metric threads
then you're in real trouble because
to a mechanic all these fasteners look
alike
and this happened on the cm1 job if i
recall we had one that the guy couldn't
figure out why it wouldn't go in the
hole
and it was a metric
course
when i had him get a gauge engage it and
we had inch stuff around there too so
and and the rest of the metric stuff was
fine thread i think and this one
happened to be of course so so this is
asking for trouble because if something
won't fit somebody's going to try to
make it fit and they put it together
now
we covered the different strength levels
and the fact that 300 series and a286
look alike
stainless steels look alike and even
different platings on materials can be
dyed to where they
they look alike
so next we go to the selection and
positioning of the washer
and you've got to pick washers that are
large enough to distribute the load
under the head or the nut
without exceeding the compressive yield
strength of the joint material
so uh
you want a hard washer and a smooth one
so that you can you know what your
coefficient of friction
is going to be
and if the internal diameter the washer
is much larger than the fastener then
you better try to try to center it
to make sure that they will fit don't do
one of those deals like i've seen people
do before where they stack up a whole
bunch of washers
and then they got to jiggle them around
to get them to fit
under the head
and you might wind up with with the
thing
embedding in the material on one side
and on the other side is hardly loaded
now shear loads on a fastener group
this is something that i i gave you
a lot of verbiage on this to help you go
through the stuff on your own
and uh so i'll just kind of hit the the
highlights on this number one on where
you have a pattern of fasteners the
first thing you want to do is determine
the centroid of the pattern but pick
picking x and y axes and using unit
areas
times its distance to get the centroid
and
although it's not a good idea to have
fasteners of a different diameter
you can
use them in this type of analysis by
ratioing the diameters
for instance the one i gave here if i
had eight bolts of a 12-volt pattern
that were three-eighths and the other
four or five-sixteenths
uh you can ratio the shank diameters and
you use
one
for the one that you have the most of
and use the the stress ratios then to
give you a factor for the other one to
use
this way you can calculate
the
the pattern
cg and get the loads on it
okay now
uh
in a lot of cases you'll have a
symmetrical pattern so you're okay
and uh
but
after you find the the centroid then you
can get these sigma r squares for the
for the fasteners which will give you an
equivalent moment of inertia if you will
like like calculating bending stresses
so
uh we can move over to the figure and i
think i can talk you through that better
uh here is a a bracket
that has a an eccentric load on it here
are
okay now to get
and it's loaded just in shear we're not
putting any tension on it
so
we have to transfer that to the cg of
course remember in strength of materials
you transfer a load
to the cg you have a
direct load and a moment is what you
replace it with
so
so you have this is your direct load is
just taking r and divide it by the
number of fasteners that gives you a
load there
now you get a moment
r times this value e
which you have to react now the way that
you react that
you take these r values
which is the distance this is the
centroid since it's a symmetrical
pattern
so you have four r values measured here
here here and here
that are the same
and then you have four more that are the
same from here to here from here to here
up to there and down to here so now you
take those and add them up so you have
four times r one squared
plus 4 times r2 squared and that gives
you your equivalent moment of inertia if
you will
then you can find a load on the fastener
by taking the moment
times
the
radius to the one that is farthest away
and that then over the sigma r squared
value that you calculated using uh those
values
and you come up with another load now
you take those two
since they're both in the shear plane
you combine them vectorially to get a
resultant load p
for a total shear load on the fastener
then of course that takes care of the
the shear loads
now if if you look at this value this
also would correspond to like a
torsional formula the tr over j in which
the sigma sigma r squared is the r sub n
squared is the equivalent of a polar
moment of inertia j
except that the load that you get here
is in pounds
and so
so then later on if we have tension on
on something like this
we can combine it
and get the total load
using stress ratios
now on edge distance and fastener
spacing this is something that's
violated a lot
in fact we put out designs around here
before
that
i have
been very disappointed with because
somebody
used practically no edge distance on
stuff we want we won't mention any names
but ron knows a guy that
did this a few times on me
but uh
here is the
edge distance and fastener spacing and
these are nominal ones so so this is
kind of what you shoot for 2d nominal
where d is the diameter of the fastener
4d spacing between fasteners
and the aircraft companies usually use a
2d plus 30 thousandths on their stuff
just to give you just a little more edge
distance in case you run into a problem
now
one of the things that questions might
be asked well if you have a shearer lug
it doesn't have 2d
no they're custom designs because
they're usually pretty thick and you go
in and calculate
hoop tension and shear tear out and that
types of things on a lug
and
that one is covered
uh i covered it in that
the chapter i wrote for that textbook
that's not out yet because that was on
on fasteners and share
and shigley and
a few other people also have uh
coverage on uh sheer and lug design
when i talk about lug you're talking
about a crank that is fairly a crank
type thing that is fairly thick so and
usually you have since it's a rotating
type
joint it does not
have a large edge distance but it's got
thick walls
now here's here's something that i
use to illustrate one of the other
fallacies that we deal with in the
engineering world
the
development of bearing stress allowables
bearing stresses
uh the the normal way of doing it you
take
this sheet is thickness t now this
represents
a uh
semicircular there would be a fastener
fitting in that hole and this represents
the the lo the way we're
coming up with the bearing stress here's
what you're actually doing because if
you have a fastener in this hole
pushing
the maximum stress is right here
so this represents that maximum stress
the here
it's zero here because you're not
putting any stress on it there
so
what we normally do and uh see uh mil
standard
1312 which we'll be covering later on in
in here gives all the different methods
of testing
of fasteners
well what they do they put the fastener
in the material and they test it to
failure
when they get it to failure whatever it
failed at for a given diameter
they divide it by the diameter times the
thickness material that's your normal
bearing area
and say that's the bearing stress so if
you look in mill handbook five or any of
these books on bearing stress allowables
you will see that they are way above
tensile element and tensile yield
because they're a fictitious thing
what they are
they're a value that has been verified
that you can use it for calculations
and get by with it but it's actually not
a true stress
so if you don't have bearing stress
allowables for material
since you see that these are
proportional p1 to the compressive yield
equals p2 to the uh
ultimate and so on you can come up with
these
just by taking one and a half times the
compressive yield or compressive
ultimate of the material
now that is a conservative figure
and because the actual test value will
run around 1.7
for most of these materials but mill
handbook 5 if they didn't test
to get the bearing allowables in a lot
of cases they'll just take one and a
half times the
tensile element or tensile yield and
slap that in there for the bearing
allowable because they know it's safe
all right grip length
and shear head and tension head on
fasteners now
grip length is a very critical thing
for shear design because that is the
length of the unthreaded portion
of the fastener
and when you're you have it in shear and
you try to keep
have no threads in the hole
so this is this is the thing that you
you do here
uh and
you're supposed to size them the
fastener such that this doesn't happen
so you put a washer under the nut
to allow tightening without running out
of threads
now the aerospace fasteners the ms nas
am that type
are available with sheer nuts or heads
or tension
heads and nuts to save weight on design
because if you're designing in shear you
don't need to have that much tension so
therefore you can go with a thinner head
or thinner nut
so we have illustrations of those in the
next figure
here is the grip length
illustration it's the bottom of the head
to the end of the threads
and
here is a shear head for a same size
fastener it's an eighth of an inch thick
down here it's uh
5 30 seconds
for a tension type
and notice two specs here that are
called out
which
those of you that are familiar with the
fasteners this is for j threads here the
mill s 8879 and this is for the two
two or class two or three and the
standard threads
here is a
shear nut and a tension nut well you see
the shear nut is pretty thin 203 versus
284 for the tension nut so if you have a
joint that is primarily shear you can
put in a little nut like that and if
you're using several hundred of them
it saves you quite a bit in weight
on a airframe
now here's here's another thing i keep
coming back to avoid tapped holes
uh we covered the tap tools and the type
of taps
and some here's some more reasons for
avoiding tapped holes cost
drilling and tapping a hole is expensive
compared to drilling a clearance hole
for a nut and bolt assembly
inspection
about the only thing you do
with a tapped hole is a go no-go gauge
and a minimum thread diameter check just
by running a pin through it
and the root radius you can't measure
very well
and since there's no such thing as a unj
tap the root radius is not rounded if
the hole's blind it'll have burrs
shavings and everything else and you're
just stuck with it
now here's the
the the other
type of design that you need to look at
is
tension loads on a fastener group
and uh
at the time that i did this one i
couldn't find one anywhere in anybody's
book so i had to draw this one up myself
but it didn't get too fatigued during
our scanning so i guess it's all right
and
the
here we have eight fasteners on a
bracket that has two different loads on
it it has a direct tension load p1 and
it has a shear load p2 which also gives
you a bending moment
so
what you're trying to do is get the
total load
on all of these fasteners
using the different loads that we have
there
all right
the
moment from the load
if you now here here's another thing
that's different from the other one is
where do you measure r from
r is measured from the healing point
for your sigma r squared
because if this thing goes into
compression over here
then you're not getting anything out of
it for your tension load so you can't
use those two fasteners to carry the
tension because they're in compression
they're going to not going to help you
any
so
what i did in this case
since it's a bracket and this is a
flange sticking out i said okay this
thing is hard up to here so i'll measure
my r's from that
uh
point to the right
so i only have for my
sigma r squared
i only have six fasteners in it
but then for the total shear
i'm using all eight of them and for the
total tension i'm using all eight of
them
so in doing that you can
calculate the sigma r squared you divide
the uh
the load i better leave this up here
from from my standpoint here for a
moment
you can divide the load by eight to get
the one the shear loads then you can
calculate
the
uh moment which is the p2 times h
then take the r7 which was the one
further stout and use the sigma r sub n
squared and you can get a load then
attention load from the moment
you have a piece of t on there which you
divide by eight to give you the
additional tension load
and you can go in then and
calculate
the total load in tension then you have
the shear load which was the p2 over
eight
and you can take those two loads now
and go in and use stress ratios
and calculate the margin of safety on
the fastener for the total loading
see here was the a better better print
showing the the one for
the uh p sub m value where you're
actually getting the moment was the p2
times h and that times r7 over the sigma
r sub n squared
now the tensile load that your preload
that you're putting on these fasteners
has to exceed p or you're in trouble
because you don't want any joint
loosening
combined shear and tension loading
now on this
you have
you get all your summation of loads in
the shear direction you get all your
summation of loads in the tension
direction
and then
you could use a mortar circle
and work with it and get the
principal stresses and that type of
thing and
calculate out
a and allowable and a margin of safety
that way but it's easier to use these
stress ratios because that's doing the
same thing
so what you do is you get two factors
you get a
r sub s or r sub t here which is the
actual shear load over the allowable
shear load for that fastener now in this
case
you can work in pounds you can work in
uh stress either when you want to as
long as you're consistent in your units
so you get a
a factor there you get one from tension
the actual tension load over the
allowable
and then you get
a margin of safety
which takes the actual load
over the one you calculated here minus
one
give a margin of safety now what happens
with these these values when you combine
them
you get
two
values
that have better be less than one for
each one of them
because you don't want either one of
them be greater than one or you're in
trouble on the design
so
you have these and they have exponents x
and y
now it depends on your degree of
conservatism as to how how big an
exponent you use for those because of
course the
the bigger the exponent goes the more
unconservative you become
because the sum of those two have to be
less than one in order to have a
positive margin
because margin of safety
and safety as a safety factor of one if
you will so a margin safety of zero is a
safety factor of one
so therefore if uh
people say oh well gee i got a margin of
safety of
0.03 on that part well that's still good
because that's 1.03 safety factor-wise
and that's the way the aerospace
industry has been doing it
ever since glenn l martin
so here are these curves that you can
use
and it depends on how conservative or
unconservative you want to be now for if
you're the belt and suspenders type and
want to make sure everything's all right
you use a straight line version here
which is
uh
just uses no exponents at all
and calculate the margin and that one is
a lot safer if you want to get
more unsafe you can go further out on
these by squaring and cubing these
ratios and that will give you
a
better margin of safety for a given load
now here's one
that is uh another one that has always
bothered me because in school
i never did
like the way these professors went
through horizontal shear stress
and said then the determination of this
is an exercise left up to the student
so
so anyway i went through and developed
this
for uh the uh lecture that i give on
fasteners
and shear
because it always bothered me that
nobody had explained it very well and of
course when you're looking for
explanations and strength materials you
go back to the basics
back to the real source timoshenko
so i found this in an old temeschenko
book
uh in which he explained it
and it was the book was old enough that
he wasn't working with bolts he was
working with nails and
tubaphores but nevertheless the
principle was the same
because when you have two pieces that
you want to fasten together so that they
act as a beam
you have to have enough fasteners to
carry the horizontal shear stress
for that to happen so this is a method
of calculating it
and this is the the vq over ib
shear stress
and
so i
set up a little problem
and work through it here and these are
the dimensions which i think most of
them are given on the
with the figure i believe
yeah
now
this is a uh
the type of beam and i i just came came
up with a kind of an artificial type
thing to illustrate the point you have a
400 pounds per inch loading
and it's 50 inches long
and it's made up of two one-inch plates
and you're wanting to hold them together
with bolts and you're going to have uh
two two at a
at each spot so you want to know how far
apart your rose bolts need to be
how far can you go and still hold the
thing together
so
that's this e is the spacing here
because you see what you actually get
when you apply the moment
then you have the horizontal shear
surface here which in this case is the
neutral axis of the beam
and you need to calculate that stress
and determine the bolts
all right
you get the reactions
to the beam
and then you get the moment
it's a uniformly loaded beam so it's wl
squared over eight
and then
you determine a value here because you
also have to check bending stress to
make sure that your bending stress is
all right even if you do carry it carry
the shear still has to hold it in
bending
so
i just took a guess at the diameter bolt
and said well i'll use a half inch bolt
in this and see how it works out and
then i'll calculate it
so if you go into this the v q over ib
remember
the
v is the vertical shear at the point
q
is the what's called the statical moment
which is the area above which you're
wanting to
check the stress above that shear plane
times
the distance to its centroid
that's a statical moment
then so so the q here was the
i calculated was was three three inches
uh cubed
because it is a
area times the distance so which makes
it cubed
then you go in and calculate the moment
of inertia in this case i left out the
diameter of the holes on this because i
was just doing it rough
and of course one of the things you that
you should do in the final calculations
you actually deduct for the diameter of
the holes in order to get the proper
moment of inertia
then i went in and said for no
bolt hole reduction
i'll have this
stress now that'll be across that shaded
area back in the figure there
which was six wide so it was six e so i
saw for
the
number of pounds that i would have that
i'd have to react at that point all
right if i take two half inch diameter
grade five bolts good for about ten
thousand five hundred pounds a piece and
i'd divide
this total load into that and solve for
e i get 2.8 inches maximum spacing
between row of bolts
so
then i went back and said okay i'll use
9 16 bolts
and with with a clearance hole and now
i'll deduct for the
holes that i'm taking out
and
calculate a new i
and then go in and calculate the
shearing stress take that to get me a
value involving e and then solve for e
using the higher allowables for the 9 16
bolts and i get 2.94 inches for row
spacing now you could optimize on that
and do all sorts of things but what i
was interested in here
was just
showing how you would do it
because uh none of the books that i had
actually strength material books showed
that the way it was supposed to be i
didn't think so
now you still have to go in and check
for beam bending and bearing stress
calculations
and
notice also that
thin structures would have to be checked
for inner rivet buckling because if you
have
thin sheet
and your fasteners are spaced too far
apart the sheet can buckle in between
fasteners under compressive load
and
i
you've heard that statement about a
little knowledge is a dangerous thing i
was
telling a guy that this could happen one
time at martin and he was taking
strength materials so he said nah
that
that can't happen
so i went to see one of the old-timers
to find out how to get out of it and he
said
well just go up and tear a page out of
his notebook that'll prove the point
because you see if you take a page
and you pull it
this way
the sheet will buckle between holes
before it tears up
so so i did that never had any more
trouble with the guy he went back to
strength materials book
so
so that's how to cut to carry the
horizontal shear loads now we go into
bolded flanges with o-rings now granted
that is a science within itself
but
we'll cover it here just to let you know
that you have to
do this with bolted joints
uh o-ring compression
in a flange is
usually just a small portion of the
total bolt load
and of course the o-ring groove is sized
to give a specific range of compression
on it when you go metal to metal on the
flanges
now for most o-rings this compression
value is like
a minimum of 10 to a maximum of about 30
percent of the unloaded cross-section
diameter
and of course the flange surfaces have
to be smooth to assure ceiling without
tearing up the o-ring
and the fastener spacing must be close
enough to keep the flanges from
separating that's one of the things you
have to watch about now granted in most
cases it's not a
problem
and of course the
this has uh just a general design
practice
you machine the o-ring groove in the
cheaper the two mating flanges
because if the machine is uh
if the machinist cuts the groove too
deep the parts scrap
so you want to make sure that it's in
the cheaper the two flanges and so you
can throw it away if you need to
and if you need a dovetail groove
to hold the o-ring in place during
assembly disassembly that also can be
machined in
now here is
a generic
o-ring joint
and uh
the uh
there's the the o-ring normally the
the only thing you got to worry about is
having a smooth enough finish on this
mating surface in that area that it
doesn't chew up the o-ring
and have enough fasteners to keep the
flanges metal to metal
now if you go to bolted flanges with
flat gaskets
then you got a another problem
you need to squeeze the gasket to seal
it
but on the other hand you don't want to
squeeze it so much
that you uh yield it in compression and
ruin it
so so now you have to look harder at the
amount of load that you're putting in
with your bolts
now a lot of gasket manufacturers would
give you a
pounds per linear inch or something for
a flat gasket so that you know them by
your bolt spacing how much you need to
put in to get the thing to seal
at least that gives you a minimum load
that you have to have and then then of
course you have to look at the
compressive yield of the gasket to see
whether you're putting too much load in
or not
so
usually the best thing to do on that is
is get the information from the
manufacturers because they they know
their their product well enough to give
you the
the proper values that you can use
and of course
we'll have some things in subsequent
sections on
what to do
where you have gaps on flanges now
here's a regular flat gasket
joint
and
one of the things you normally do with
gaskets too if you're in
the automotive world you use some sort
of a gasket cement sealer or something
of that nature on them
to
stick them in place
while you're putting the joint together
and we have
loading curves for the
flat gaskets in the appendix which you
get one of these days here in the near
future
and
the
flat gasket joint design
bickford has quite a bit more coverage
on it
and
so between that and the manufacturers
chances are you can come up with enough
information for that
now gasket loads in flange joints
leaks usually start at the point of
maximum flange bending which is midway
between adjacent uh bolts where the
gasket's not compressed enough to seal
uh a lot of you have run into that in
the past with
valve covers on cars
they don't have enough
fasteners in them and you have cork
gaskets so you tighten them down and the
thing will bow
and leak in the middle and so you have
you put cardboard under your car
and so to increase the load at the
midway point you can look at three
different ways of doing it one is to
increase the number of bolts
increase the flange thickness
and increase the initial bolt torque
so those are three things that you can
look at all of which have their
advantages and disadvantages
the increased number of bolts
since deflection is proportional to the
cube of the span between bolt centers
that cuts way down on the deflection of
the flange
and so uh
so adding a bolted mid stand gives you
gives you a
cut of 8 on the
deflection
and increases the gasket load the only
thing is
you increase the cost because you've now
added another bolt
another bull hole
increasing the flange thickness
now since flange deflection is inversely
proportional the cube to the flange
thickness you double the thickness it
decreases the flange deflection by a
factor of eight
so
that is a good thing except that
you increase the weight and the cost of
material so that's another thing that
you have to weigh
now increasing the bolt torque
is
the cheapest way of doing it
but if you increase it to a certain
point
the flange can bend
in the middle
because you're compressing it down under
the bolt and allow it to bow up in the
middle where to leak worse so
particularly if you have a soft gasket
like the kirk gaskets
you you get leakage
so and if the bolt is near the yield
point a further increase in torque can't
be made unless you use bolts with a
higher strength
so one of the el cheapo ways that you
can do on this
is put
extra diameter type uh washers under the
bolts to spread the load out just a
little bit sometimes that'll stop them
from leaking but that's not something
you'd want to put in an original design
now getting
into bolted flanges for glass windows
the reason i put this in is this is a
special one and we've used it around
here on
designing of windows for
pressure vessels because normally you
don't think of a window as
needing
much in the way of gaskets and they more
or less just slap them in
and they're done with it they're in
cameras and things of this nature but
where you need to
use
sight gauges
or in our case actually you in cm1 use
cameras through glass
now you wind up with an optical quality
window
that costs
several thousand dollars
that you need to make sure that nothing
happens to it
so the way to make sure nothing happens
to it is that you kind of pat it all the
way around with rubber
to keep it from touching the metal and
then of course the
one of the things about it in ours in
particular was that the
fastener design becomes a balancing act
to seal it without overloading it
because you don't want to overload it
and the other thing too is glass is so
brittle
the thing that causes it to fail of
course is surface imperfections and
you can't afford to scratch it with
anything
so uh it has a coefficient of thermal
expansion about 1 6 that of metal so if
you put it in and you have a temperature
change
now you have to put enough
padding around it of some sort
to allow it
to be compressed
uh by the
material around it or expanded and so on
without leaking so that's a balancing
act
so you sandwich it
in flat rubber gaskets with a bumper
strip around the outside of the window
to keep it from direct contact with the
metal
and
then
the balancing act is to seal it so that
it won't leak
but yet not
compress it too much
so we have a typical
design
shown in figure 35 which you can go and
put up there
and
now
this is a
model of what we actually used in cm1
except
that the one thing that i didn't show
just for clarity was the o-ring that we
had there but you see here i'll i'll use
this one it's a little clearer
here is the window
you have a bumper strip around the
outside now this is something that's not
a sealer it's just to keep it when you
drop it in the socket that it's in or
the well there to keep from touching it
you have rubber gaskets on the bottom
you have a round gasket there
and then you have one on top
now what what you're doing here
you're going metal to metal with this
top flange
now you have to use the tolerances of
both
on machining of this
and machining of this surface in order
to
make sure that you can put that a window
in there
size your gaskets properly in some cases
you have to grind them to get them to
the right diameter i mean right
thickness because the rubber is not
close enough tolerance
put it in there bolt it all down
and seal it
without hurting anything
so that is a special design within
itself and of course with your bolts you
have to make sure that they're strong
enough to
go metal to metal and load the thing up
without over stressing bolts
now the effect of friction in a clamp
joint
in most cases
friction forces
between
clamp surfaces
are not included in the shear
calculations
the reason being
that's too hard to determine what they
are
because if you've got
oil or grease on the surfaces
the your friction coefficient could be
real low
if you've got
striations of some kind on it it could
get real high but you really don't know
what it is
so for that reason you normally
don't include
the
axial force of the bolt times the
coefficient of friction
as a
shear
capability that you have now there are a
few cases and the next page will show
there are the actual friction forces
that you can get
you see you have a bolt preload of p
and so you take that times the
coefficient of friction on this surface
and this surface and you got two forces
here
that
two
of these n forces and the n is the
friction load
uh which is the normal load
times the coefficient of friction by
definition
so
in some cases in the construction world
they actually
use
this friction
load
when they
are doing the joint calculations because
they count on it
but you it's not something that you
would normally count on because it's too
unpredictable
now the
compression
cone of a bolted joint
we covered this earlier there
in the stiffness section
but uh
in the appendices we do give more stuff
on it and the
here's something that i alluded to
earlier the bulk joint relative
stiffness calculations
and
most of the ordinary designs are not a
big requirement it's just that where you
have say small areas
that you would want to uh like for
instance if you had some now if uh one
of the things that would be a real
problem if you have a bushing type thing
around the bolt or a spacer or something
like that then then you better go in and
check happen real fast because you could
get into trouble but if you have
uh
say
uh steel and you're using steel bolts
chances are the joint is going to be
stiff enough that you don't have a
problem with it you can look at it
and take the the method at least work on
it calculate a circular
model for the stiffness if that is
satisfactory then go no further now if
if you were bowling say through
all
soft aluminum copper something like that
then you would probably want to do some
joint stiffness calculations to make
sure that you're not in trouble
but in in most cases you can get by with
a minimal amount of joint stiffness
calculations
now bolding of dissimilar materials
as i mentioned earlier in the centaur
case where you went from room
temperature to -300 or something like
that
with dissimilar materials there you got
a real problem because of the
differential thermal expansion and
contraction
of the materials because aluminum is
let's see something like
three times i believe isn't it the uh
on thermal
something like three times as
high a coefficient of steel is
and copper is way up there so if you're
if you're bolding up a copper joint you
have a temperature change you got to be
real careful on it
but uh
then the the other thing
that you needed watch for is the
galvanic corrosion
because unless the mating surfaces are
insulated from each other
and that was one of the reasons
why the
magnesium is kind of going out of vogue
because how do you
how do you
insulate it satisfactorily that over a
period of 20 years if it's used in the
airplane component that it's going to
stay insulated
because a lot of these
organic type things that they use for
insulation
the sealers that they put around rivets
bolts and stuff like that on airplanes
over a period of years can deteriorate
and moisture and the other the other
problem with the
the fasteners on an airplane is that
most of them
you're looking at heads
sticking out so if you have a crack that
is starting at the edge of the hole
it has to come out quite a ways before
you can see it so so that has caused
a lot of problems there
the other thing that you need to look at
is the yielding of softer materials
because if you are say using a
high-strength bolt in aluminum and you
crank that up too much you can actually
yield the aluminum in compression under
the head
without doing anything to the bolt
so uh
and then of course you've got to check
the strengths at the temperature
extremes because like for instance
aluminum
uh falls off drastically
at only 250 degrees whereas steel most
steels will go up to 700 and up before
they start falling off in strength so so
if you were
tightening up an aluminum joint and you
ran it up to say 250 300 degrees you
could get yielding real easy under the
heads of the bolts
now maximizing the effective length of
fasteners of course when we discussed
the stiffness ratios the effective
length of the fastener was mentioned
and this is important on the
differential expansion contraction
so
it may be necessary to add a spring
or a belleville washer under a bolt head
to increase its effective length enough
to satisfy the design so that it won't
loosen up and
the deflection of course is the pl over
e so you increase l you're
doing all right on it in fact on the
exhaust system
on uh
some of the ford trucks
they actually have a big spring
on the bolt that holds the flange to the
catalytic converter
it's put on there i think to take the
temperature
differential that you get between the
materials
on because you can go from room
temperature up to about
1300 degrees or something like that on
them
okay we'll take a break cure for now and
come back up with the match drilling of
fastener holes
which is an important one