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18. Threaded Fasteners - University of Arizona Flipbook PDF
J. H. Burge University of Arizona 3 Pitch/diameter charts Note: To determine the tap drill size for metric fasteners, si
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18. Threaded Fasteners We use threaded fasteners (screws) to hold things together. How are these specified? What are the issues when choosing among different types of screws?
How to specify fasteners Inch Dia. (inch)
Threads per inch
Length Drive (inch) Style
Head Style
Strength Level
Plating/ Coating
3/4 -
10
X
5
Hex
Cap Screw
Grade 5
Zinc
3/4 -
16
X
5
Hex
Cap Screw
Grade 5
Zinc
Metric Dia. (mm)
Pitch (only if fine)
M8
X
M8 X
1
J. H. Burge
Length Drive (mm) Style
X
Head Style
Strength Level
25
Hex
Cap Screw
8.8
25
Hex
Cap Screw
8.8
University of Arizona
DIN No.
Plating/ Coating
931 Plain (partial thread) 933 (full thread) 960 Plain (partial thread) 961 (full thread)
1
Thread pitch
Inch
Metric
Number of threads per inch
Distance between crests of adjacent threads
Diameter inches
3/4 3/4
-
J. H. Burge
Threads per inch
Type
Diameter millimeters
10 16
= CoarseUNC = Fine UNF
M8 M8
University of Arizona
Distance millimeters
X X
1.25 1
Type
= Coarse = Fine
2
Pitch/diameter charts
Note: To determine the tap drill size for metric fasteners, simply subtract thread pitch from the fastener diameter and drop all but the first decimal place. Example: M12 - 1.75 = 10.25 is a 10.2 tap drill size
J. H. Burge
University of Arizona
3
Drills and taps for common threads Major diam. (inches)
Clearance Drill
UNC tpi
Tap Drill for UNC
UNF tpi
Tap Drill for UNF
Nut Size
0
0.0600
#52
—
—
80
3⁄ ″ 64
5⁄ ″ 32
2
0.0860
#43
56
#50
64
#50
3⁄ ″ 16
4
0.1120
#32
40
#43
48
#42
¼″
6
0.1380
#27
32
#36
40
#33
5⁄ ″ 16
8
0.1640
#18
32
#29
36
#29
11⁄ ″ 32
10
0.190
#9
24
#25
32
#21
3⁄ ″ 8
¼″
0.2500
F
20
#7
28
#3
7⁄ ″ 16
5⁄ ″ 16
0.3125
P
18
F
24
I
9⁄ ″ 16
3⁄ ″ 8
0.375
W
16
5⁄ ″ 16
24
Q
5⁄ ″ 8
7⁄ ″ 16
0.4375
29⁄64″
14
U
20
25⁄ ″ 64
½″
0.5000
33/64″
13
27⁄ ″ 64
20
29⁄ ″ 64
9⁄ ″ 16
0.5625
9⁄ ″ 16
12
31⁄ ″ 64
18
33⁄ ″ 64
5⁄ ″ 8
0.6250
5⁄ ″ 8
11
17⁄ ″ 32
18
37⁄ ″ 64
¾″
0.7500
¾
10
21⁄ ″ 32
16
11⁄8″
7⁄ ″ 8
0.8750
7⁄ ″ 8
9
49⁄ ″ 64
14
15⁄16″
1″
1.0000
1″
8
7⁄ ″ 8
14
1½″
Gage and Fractional Sizes
J. H. Burge
University of Arizona
¾″
4
Common head styles
Drive Types
Phillips and Frearson An X shaped drive. Abbreviated PH
Slotted A slot in the head. Abbreviated SL
Combination A combination of slotted and Phillips drives. Abbreviated Combo
Socket or Allen A hexagonal hole for use with and Allen wrench.
(Ref McMaster-Carr)
J. H. Burge
University of Arizona
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Socket head cap screw basics
Socket head cap screw
Low head
Flat head
Button head
Socket shoulder screw
Socket Head Cap Screw - strongest of all head style. • •
Head height is equal to shank diameter. Shouldn't be mated with a regular hex nut, which isn't as strong.
Low Head Cap Screw - designed for applications with head height limitations •
Head height is approximately half the shank diameter.
Flat Head Cap Screw - for flush applications Caution: Inch and metric have different countersink angles. Mismatching fastener and hole countersink angles can result in premature fastener failure
Button Head Cap Screw • • •
Larger head diameter makes it more appropriate for holding thin materials like sheet metal guards. Because of its internal hex drive style it's ideal for tamper-proofing applications. Good substitute for other drive styles that are prone to stripping like Phillips and slotted.
Socket Shoulder Screw •
Typically used as a pivot point or axle because shoulders are ground to a tight tolerance.
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University of Arizona
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Thread Class • •
Socket head cap screws 1" diameter and under have class 3A fit (tight tolerance). Over 1" have a class 2A fit (typical hex cap tolerance). All other head styles and diameters have a class 3A fit.
Specifications
Head and Body Dimensions - Alloy and Stainless - ASME B18.3-1998 - Includes 1960 Series D Screw Diam. 0 1 2 3 4 5 6 8 10 1/4 5/16 3/8 7/16 1/2 5/8 3/4 7/8 1 1-1/4 1-1/2
Body Diameter .0600 .0568 .0730 .0695 .0860 .0822 .0990 .0949 .1120 .1075 .1250 .1202 .138 .1329 .1640 .1585 .1900 .1840 .2500 .2435 .3125 .3053 .3750 .3678 .4375 .4294 .5000 .4919 .6250 .6163 .7500 .7406 .8750 .8647 1.0000 .9886 1.2500 1.2336 1.5000 1.4818
J. H. Burge
A
H
C J F T Top Hex Fillet Key Chamfer Head Head Socket Transition Engageor Diameter Height Size Diameter ment Radius .0960 .091 .060 .057 .004 .050 .074 .063 .025 .118 .112 .073 .070 .005 1/16 .087 .076 .031 .140 .134 .086 .083 .008 5/64 .102 .090 .038 .161 .154 .099 .095 .008 5/64 .115 .103 .044 .183 .176 .112 .108 .009 3/32 .130 .118 .051 .205 .198 .125 .121 .012 3/32 .145 .132 .057 .226 .218 .138 .134 .013 7/64 .158 .145 .064 .270 .262 .164 .159 .014 9/64 .188 .173 .077 .312 .303 .190 .185 .018 5/32 .218 .202 .090 .375 .365 .250 .244 .025 3/16 .278 .262 .120 .469 .457 .312 .306 .033 1/4 .347 .329 .151 .562 .550 .375 .368 .040 4/15 .415 .398 .182 .656 .642 .438 .430 .047 3/8 .484 .465 .213 .750 .735 .500 .492 .055 3/8 .552 .532 .245 .938 .921 .625 .616 .070 1/2 .689 .664 .307 1.125 1.107 .750 .740 .085 5/8 .828 .801 .370 1.312 1.293 .875 .864 .100 3/4 .963 .933 .432 1.500 1.479 1.000 .988 .114 3/4 1.100 1.069 .495 1.875 1.852 1.250 1.236 .144 7/8 1.370 1.334 .620 2.250 2.224 1.500 1.485 .176 1 1.640 1.602 .745
University of Arizona
G
K Bottom Wall Chamfer Thickor ness Radius .020 .007 .025 .007 .029 .007 .034 .007 .038 .008 .043 .008 .047 .008 .056 .008 .065 .008 .095 .010 .119 .010 .143 .010 .166 .015 .190 .015 .238 .015 .285 .015 .333 .020 .380 .020 .475 .020 .570 .020
7
Head Dimensions Hex Head and Hex Washer Head Machine Screws - ANSI B18.6.3 A
W
H
Width Across Flats
Width Across Corners
Head Height
Max Min
Min
Max Min Max Min Max Min Max Min Max Min
4
.188 .181
.202
.060 .049 .243 .225 .019 .011 .039 .031 .042 .025
5
.188 .181
.202
.070 .058 .260 .240 .025 .015 .043 .035 .049 .030
6
.250 .244
.272
.093 .080 .328 .302 .025 .015 .048 .039 .053 .033
8
.250 .244
.272
.110 .096 .348 .322 .031 .019 .054 .045 .074 .052
10
.312 .305
.340
.120 .105 .414 .384 .031 .019 .060 .050 .080 .057
12
.312 .305
.340
.155 .139 .432 .398 .039 .022 .067 .056 .103 .077
1/4
.375 .367
.409
.190 .172 .520 .480 .050 .030 .075 .064 .111 .083
5/16
.500 .489
.545
.230 .208 .676 .624 .055 .035 .084 .072 .134 .100
3/8
.562 .551
.614
.295 .270 .780 .720 .063 .037 .094 .081 .168 .131
Nominal Size
J. H. Burge
F
U
Washer Washer Diameter Thickness
University of Arizona
J
T
Slot Width
Slot Depth
8
Strength of fasteners
1 Amount of force required to pull apart fastener 2 Converted from megapascals (MPa) PSI/145 = MPa or MPa x 145 = PSI
Inch • •
Grade indicates strength level in the inch system. When specifying inch fastener strengths, call them out as grades.
Metric • •
Property class indicates strength level in the metric system. When specifying metric fastener strengths, call them out as property classes. Don't confusingly ask for a metric Grade 8 hex cap screw. Ask for a metric property class 10.9 instead.
Warning: Similar numbers used for grade and property class designations don't stand for the same strength. For example, a Grade 8 hex cap screw and a property class 8.8 hex cap screw have different tensile strengths. See table.
J. H. Burge
University of Arizona
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Strength of threads Shear stress is total force/engaged area Rules of thumb: Engage screws into threads over length 1.5 x the diameter Root diameter = screw diameter – thread spacing Shear strength = ultimate strength/sqrt(3) (using Von Mises strength) Example: ¼-20 grade 2 screw threaded into Aluminum For 1000 lb load Strength of Al threads For engaged length L = 0.37 in Root diameter Dr= ¼ -(1/20) = 0.2” Engaged area = pi * D L = 0.24 in2 Shear stress = 4250 psi Ultimate strength of aluminum= 40 ksi Shear strength = 40 ksi/1.73 = 23 ksi. Safety factor of 23/4.3 = 5.4
Strength of screw: Root diameter Dr = 0.25 – 1/20 = 0.2” A = pi Dr2/4 = 0.031 in2 Stress = 1000 lb/0.031=31ksi Ultimate strength for grade 2 bolt is 74 ksi Proof load strength is 55 ksi Safety factor of 55/31 = 1.8
More detailed method of establishing strength is given in the appendix Threaded inserts: Threads in soft materials are easily damaged Strength can be significantly improved
J. H. Burge
University of Arizona
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About Socket Cap Screw Materials Finish/Coating
Features
Plain
Good for general purpose applications.
Zinc-Plated
Provides excellent corrosion resistance.
Cadmium-Plated
Offers better rust resistance than zinc-plating, especially in salt environments.
Nickel-Chrome Plated
Polished and buffed to a bright, mirror-like finish. Resists wear and corrosion.
Black-Oxide
Offers mild rust resistance and some lubrication qualities.
Blue-Coated
This highly visible blue coating makes it easier to distinguish metric from inch sizes.
Ultra CorrosionResistant Coated
Also known as armor coat. Provides better corrosion resistance than zinc, cadmium, and hot-dipped galvanized plating. The thickness of the coating does not interfere with the thread fit.
Material Type
Features
Plain Steel
Good for general purpose applications.
18-8 Stainless Steel
Provides excellent corrosion resistance. May be mildly magnetic.
300 Series Stainless Steel
Meet more stringent specifications such as military specifications. Corrosion Resistant.
316 Stainless Steel
Offers excellent corrosion resistance, even more than 18-8 stainless steel. Contains molybdenum which increases corrosion resistance to chlorides and phosphates.
Bumax 88 Stainless Steel
316L stainless steel with a high molybdenum content offering corrosion resistance similar to 316 stainless steel. May be mildly magnetic.
Brass
Nonmagnetic and softer than stainless steel and mild steel.
Nylon 6/6
Nonconductive and resistant to chemicals and solvents (except mineral acids). Since nylon absorbs moisture from the environment, changes in moisture content will affect the fastener's dimensions and properties. Withstands a wide range of temperatures.
Silicon Bronze
Made of 95-98% copper with a small amount of silicon for strength. Nonmagnetic and offers high thermal conductivity and corrosion resistance.
A286 Super Alloy
Made of 26% nickel and 15% chrome with corrosion resistance similar to 18-8 stainless steel and strength properties comparable to alloy steel. Is considered an iron-based super alloy. Passivated (a nitric acid treatment that creates a passive film to protect against oxidation and corrosion).
.
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University of Arizona
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Washers • • • • • •
Distribute load from screw head Protect surface from screw head Keep screw from backing out Take up space (shim) Act as a spring Provide sealing
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University of Arizona
12
Bolted joints: Stiffness comes from assembly
J. H. Burge
University of Arizona
13
Suggested Tightening Torque Values to Produce Corresponding Bolt Clamping Loads SAE Grade 2 Bolts
SAE Grade 5 Bolts
SAE Grade 8 bolts
74 ksi tensile strength
120 ksi tensile strength
150 ksi tensile strength
55 ksi proof load
85 ksi proof load
120 ksi proof load
Bolt
Stress
Clamp
Torque
Torque
Clamp
Torque
Torque
Clamp
Torque
Torque
Diam.
Area
Load
Dry
Lubed
Load
Dry
Lubed
Load
Dry
Lubed
Size
D(in.)
A(in²)
P (lb)
in-lb
in-lb
P (lb)
In-lb
in-lb
P (lb)
in-lb
in-lb
4-40
0.1120
.00604
240
5
4
380
8
6
540
12
9
4-48
0.1120
.00661
280
6
5
420
9
7
600
13
10
6-32
0.1380
.00909
380
10
8
580
16
12
820
23
17
6-40
0.1380
.01015
420
12
9
640
18
13
920
25
19
8-32
0.1640
.01400
580
19
14
900
30
22
1260
41
31
8-36
0.1640
.01474
600
20
15
940
31
23
1320
43
32
10-24
0.1900
.01750
720
27
21
1120
43
32
1580
60
45
10-32
0.1900
.02000
820
31
23
1285
49
36
1800
68
51
1/4-20
0.2500
0.0318
1320
66
49
2020
96
75
2860
144
108
1/4-28
0.2500
0.0364
1500
76
56
2320
120
86
3280
168
120
5/16-18
0.3125
0.0524
2160
11
8
3340
17
13
4720
25
18
5/16-24
0.3125
0.0580
2400
12
9
3700
19
14
5220
25
20
3/8-16
0.3750
0.0775
3200
20
15
4940
30
23
7000
45
35
3/8-24
0.3750
0.0878
3620
23
17
5600
35
25
7900
50
35
7/16-14
0.4375
0.1063
4380
30
24
6800
50
35
9550
70
55
7/16-20
0.4375
0.1187
4900
35
25
7550
55
40
10700
80
60
1/2-13
0.5000
0.1419
5840
50
35
9050
75
55
12750
110
80
1/2-13
0.5000
0.1599
6600
55
40
10700
90
65
14400
120
90
9/16-12
0.5625
0.1820
7500
70
55
11600
110
80
16400
150
110
9/16-18
0.5625
0.2030
8400
80
60
12950
120
90
18250
170
130
5/8-11
0.6250
0.2260
9300
100
75
14400
150
110
20350
220
170
5/8-18
0.6250
0.2560
10600
110
85
16300
170
130
23000
240
180
3/4-10
0.7500
0.3340
13800
175
130
21300
260
200
30100
380
280
3/4-16
0.7500
0.3730
15400
195
145
23800
300
220
33600
420
320
Notes: 1. Tightening torque values are calculated from the formula T = KDP, where T= tightening torque. lb-in. K=torque-friction coefficient; D = nominal bolt diameter. in; and P = bolt clamp load developed by tightening. lb. 2. Clamp load is also known as preload or initial load in tension on bolt. Clamp load (lb) is calculated by arbitrarily assuming usable bolt strength is 75% of bolt proof load(psi) times tensile stress area(sq in.) of threaded section of each bolt size. Higher or lower values of clamp load can be used depending on the application requirements and the judgement of the designer.
J. H. Burge
University of Arizona
14
Appendix Guide to Specifying Torque Values for Fasteners Note : The following notes are given as a guide only. It is recommended that torque values derived from formulae should not be used without comparison to figures obtained using practical tests. Introduction Generally, in the majority of applications, the reliability of the joint is dependent upon the bolt's ability to clamp the parts together. Adequate clamping prevents relative motion between parts of the joint and leakage from joints containing gaskets. Measuring a bolt's clamp force is difficult, especially under production assembly conditions. The clamp force generated by a bolt can be indirectly controlled by regulating the applied torque. The method, known as Torque Control, is by far the most popular method of controlling a bolt's clamp force. The initial clamp force generated by the bolt is frequently called Preload. There is a link between the torque applied to a bolt and the resulting preload. A problem exists because friction has a large influence on how much torque is converted into preload. Besides the torque required to stretch the bolt, torque is also required to overcome friction in the threads and under the nut face. Typically, only 10% to 15% of the torque is used to stretch the bolt. Of the remaining torque, typically 30% is dissipated in the threads and 50% to 55% under the nut face. Because friction is such an important factor in the relationship between torque and preload, variations in friction have a significant influence on the bolt's preload. Different bolt surface finishes generally have different friction values. The torque required for a socket headed screw will not be the same as that required for the same size hexagon bolt. The larger bearing face of the standard bolt will result in increased torque being required compared to a socket headed screw. This is because more torque is being dissipated between the nut face and the joint surface. Stresses induced into a bolt When a bolt is tightened, the shank and thread sustain a direct (tensile) stress due to it being stretched. In addition, a torsion stress is induced due to the torque acting on the threads. These two stresses are combined into a single equivalent stress to allow a comparison to be made to the bolt's yield strength. In order to effectively utilize the strength of the bolt, yet leave some margin for any loading the bolt would sustain in service, an equivalent stress of 90% of yield is commonly used. This approach is used in this guide. This approach has a number of advantages over the method where a direct stress, and hence preload value, is assumed in the bolt. For high thread friction values, a high torsion stress results in the bolt. Less of the available strength of the bolt is being utilized in such a case to generate preload. In the extreme case when a nut has seized on the bolt thread, all the applied torque is sustained as torsion stress
J. H. Burge
University of Arizona
15
with no preload being available. In the other extreme, low thread friction results in higher preloads. Note : The following information is provided to assist Engineers wishing to establish the theoretical torque value for a particular fastener. Caution should be exercised when using theoretical values because the preload and torque is dependant upon the friction values selected. Calculation Procedure The formulae used are applicable to metric and unified thread forms which have a thread flank angle of 60o. The calculation procedure distinguishes between thread and underhead friction as well as differences which can be caused by bearing face diameter variations. The procedure comprises of the following steps; 1. Fastener Details Dimensions and strength grades are specified in various standards.
Table 1 Strength Grade
3.6 4.6 4.8 5.6 5.8 6.8 2
* Yield Stress N/mm
8.8
9.8 10.9 12.9
180 240 320 300 400 480 640 # 720 900 1080
* Nominal values quoted. # For grades 8.8 and above a proof stress is specified because of problems measuring yield. BS 6104 Pt. 1 Table 1 presents information on strength grades of bolts; the most common grade for metric fasteners is grade 8.8. Estimating the appropriate friction coefficient can problematic.
Table 2 External Steel Threads
Internal Self Finish Steel Threads
Internal Zinc Plated Steel Threads
Internal Cast Iron Threads
Internal Aluminium Threads
Dry Self Finish or Phosphate Treated
0.10 to 0.16
0.12 to 0.18
0.10 to 0.16
0.10 to 0.20
Oiled Self Finish or Phosphate Treated
0.08 to 0.16
0.10 to 0.18
0.08 to 0.18
0.10 to 0.18
Dry Zinc Plated
0.12 to 0.20
0.12 to 0.22
0.10 to 0.17
0.12 to 0.20
Oiled Zinc Plated
0.10 to 0.18
0.10 to 0.18
0.10 to 0.16
0.10 to 0.18
Thread Adhesive
0.18 to 0.24
0.18 to 0.24
0.18 to 0.24
0.18 to 0.24
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University of Arizona
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Tables 2 and 3 may be used as a guide when other information is not available.
Table 3 Condition of the Bolt Head or Nut
Zinc Plated Steel part clamped by Bolt
Self Finish Steel part clamped by Bolt
Cast Iron Aluminum part part clamped clamped by by Bolt Bolt
Dry Zinc Plated Finish
0.16 to 0.22
0.10 to 0.20
0.10 to 0.20
-
Slight Oil Applied to Zinc Plated Finish
0.10 to 0.18
0.10 to 0.18
0.10 to 0.18
-
Dry Self Finish or Phosphate or Black Oxide Finish
0.10 to 0.18
0.10 to 0.18
0.08 to 0.16
-
Slight Oil Applied to a Self Finish or Phosphate or Black Oxide Finish
0.10 to 0.18
0.10 to 0.18
0.12 to 0.20
0.08 to 0.20
Gaps in table indicate a lack of available published data. 2. Determination of the tensile stress in the threaded section. To determine the tensile stress in the fastener, first establish what proportion of the yield strength you wish the tightening process to utilise. Normally a figure of 90% is acceptable but may be varied to suit the application. Because of the torque being applied to the threads, torsion reduces the tensile stress available to generate preload. The following formula can be used to determine the tensile stress in the thread.
3. Establish the preload The preload F is related to the direct tensile stress
by :
The stress area of the thread As represents the effective section of the thread. It is based upon the mean of the thread pitch and minor diameters. It can be obtained from tables or calculated using the formula:
J. H. Burge
University of Arizona
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4. Determine the tightening torque. The relationship between tightening torque T and bolt preload F is:
If units of Newton's and millimeters are being used, T will be in N.mm. To convert to N.m, divide the value by 1000. The effective friction diameter Df can be determined using the following formula:
For a standard hexagon headed nut, Do is usually taken as the across flats dimension and Di as the diameter of bolts clearance hole. Note : Use of friction values As can be seen from tables 2 and 3, upper and lower limits to friction values are stated. Traditionally a mean value of friction is used when calculating the tightening torque and preload value. Be aware however, that for other conditions remaining constant, the higher the value of friction - higher is the required tightening torque and lower is the resulting preload.
Terms used in the formulae T
Tightening torque to be applied to the fastener.
F
The preload (or clamp force) in the fastener. Equivalent stress (combined tensile and torsion stress) in the bolt thread. A figure of 90% of the yield of proof stress of the fastener is usual. Tensile stress in the fastener.
J. H. Burge
d2
Pitch diameter of the thread.
d3
Minor (or root) diameter of the thread.
P
Pitch of the thread.
µT
Thread friction coefficient.
µH
Friction coefficient between the joint and nut face.
Df
The effective friction diameter of the bolt head or nut.
D0
Outside diameter of the nut bearing surface.
Di
Inside diameter of the nut bearing surface.
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Example Calculation As an example, the above formulae will be used to determine the preload and tightening torque for a grade 8.8 M16 hexagon headed bolt. Step 1 Establishing the dimensions and friction conditions. The data below is to be used. d2 = 14.701 mm d3 = 13.546 mm P = 2 mm µT Taken as 0.11 µH Taken as 0.16 Step 2 Calculating the tensile stress in the fastener using 90% of 640 N/mm2 gives 576 N/mm2, substituting values into the formula gives;
=
= 491 N/mm2. Step 3 Taking the stress area as As as 157 mm2, gives the bolt preload F to be 77087N. Step 4 Determination of the tightening torque T. i ) The effective friction diameter. Taking D0 = 24 mm and Di = 17.27 mm gives Df = 20.6 mm. ii ) Using the values calculated gives a tightening torque T of 223481 , that is 223 Nm.
J. H. Burge
University of Arizona
19