introduction, drawing, calculation for winch design

1.0

INTRODUCTION
A winch is a mechanical device that is used to pull in (wind up) or let out (wind out) or

otherwise adjust the "tension" of a rope or wire rope (also called "cable" or "wire cable"). In its
simplest form it consists of a spool and attached hand crank.
In larger forms, winches stand at the heart of machines as diverse as ;
-

Tow trucks, steam shovels and elevators.

-

The spool can also be called the winch drum.

-

Can be powered by electric, hydraulic, pneumatic or internal combustion drives.

-

Some may include a solenoid brake or a mechanical brake or ratchet and pawl device that
prevents it from unwinding unless the pawl is retracted.

In the operation system, it can be conducted by using convention or non-convention system.
Convention system consist of mainly human effort or force distributed by human and the cost of
operation is cheaper because it does not consist of high tech equipment and also low in
maintenance.
Non-convention system is for most technologies that we see today which the entire burden is
been carried out by a machine. Efficiencies rate are also greater then convention system but the
maintenance can be high due to the system it have.
Example that we can take is such as elevator that can carried passengers more than 1 ton
(1000 kg) and the system is more into non-conventional because it uses motorized engine to
carry passengers. Other example that we can take is like steam shovel. Used in lifting and
moving material such as rock and soil. It is the earliest type of power shovel or excavator.
For our project, we are required to design a winch according to the specification given.
The first requirement is to design a winch that can withstand maximum load of 15kN. The
second requirement is we need to use cable of 14mm diameter.

2.0

WINCH

EMM 3506 – GROUP 3

Page 1

2.1

PROBLEM STATEMENT

A boat without an anchor is like a kite without thread. The boat will be drifting due to the
wind and wave. So, having an anchor is a must for every boat. If using bare hand to pull up the
anchor, it will make the job tiring and complicated. In order to overcome the problem, they will
need a very useful mechanical device to help them pulling up the anchor to the boat and to
unload from the boat.
The very best way to overcome this problem is by installing a winch to the boat.
Winch was widely used to pull in ( wind up) or let out ( wind out) or otherwise adjust the
tension of a rope either it wire or nylon, but in this case, we used wire cable. In its simplest form,
it is consists of a pool ad attached hand crank that can help and overcomes those:
o
o
o
o

Wear on boat’s body which is lead to corrosion
Conserve human energy
Keep the wire cable last longer
Wire cable stored in good manner and keeps the safety environment on the boat

That problem that we need to overcome and to achieve is the main reason why the
application of winch is suitable for anchor problem, which is suitable case for our product to be
used widely in this sector.

Figure 1: problem overview scope
2.2

FEATURES AND APPLICATION


Page 2

As a project that must be conduct by a group of student, the requirement for a winch to be
consider as a winch in real words was be given out. Refer on the design factors that already be
listed and was mention in characteristic of the winch, we, group 3 come out with rough features
and the application for the winch based on design factors listed below:
Design Factors
o
o
o
o
o
o
o

Economic, environmental and safety toward users
Ergonomic
Ease to assembly
Ease of fabrication
Energy efficiency
Maintenance requirements
Environmentally friendly

Components
o
o
o
o
o
o
o

Wire rope ( 14mm in diameter )
Drum
Ratchet wheel and Pawl
Bearing
Frame
Gear / speed ratio
Brake

Features
7. Stainless steel drive shaft and ball bearing.
8. Lightweight stainless steel drum with large diameter to reduce wire rope wear.
9. Ergonomic handle for easy and good grip for handling.
10. Drum is located horizontal to ensure less wire rope wear.
Applications
 Anchor pulley

2.3

DESIGN REQUIREMENT

Corrosion-resistant is must in this case study, which is corrosion easily happen due to sea water.

Page 3

For easy handling during wind up and out, this design come out with 2 speeds for easy
control and constant power required for anchor to be pulled. However, the neutral gear also been
consider for safety factor. Freewheeling position for fast line payout.
All these requirement and features are great for marine application precisely.
Features:
o Freewheeling winch is ideal for boat trailers.
o 2 speed to fit application in marine
 Fast speed
– lets you have full speed to pull up anchor quickly.
 Lower speed – offers an increase in mechanical advantages.
o New features, shift lock design – lets you change gears without having to pull out the
handle.
 The mechanism based on slide along the shaft.
 Hold the shaft in gears position.
o Drum – the drum is design to withstand load up to 15kN and can store about 50m in
length of wire cable.
o Wire cable – the selection of wire are been consider for marine used and can be withstand
natural phenomena.

3.0

COMPONENTS


Page 4

3.1

WIRE ROPE SPECIFICATION

Figure 2: wire selection table

Length of rope: 50000 mm
Nominal breaking load : 133 kN
Weight: 0.86 kg/m
Rope construction: 6 x 26 RRL(right regular lay) rope
Safety factor of rope =

(
(

)

)

=
=8.87

3.2

HOOK SPECIFICATION


Page 5

Figure 3: table of selection of hook

4.0

DRUM

4.1

CALCULATIONS


Page 6

Parameter required




Maximum load
Diameter of rope
Length of rope

=15 kN
= 14 mm
= 50000mm

Calculation
1. Diameter of drum
Ddrum = (ratio between 20 to 25) x drope
=20 x 14
=280 mm
2. Groove radius,
r

3. Groove diameter,
d

= 0.53 x d

= groove radius x 2

= 0.53 x 14mm

= 7.4mm x 2

= 7.4mm

= 14.8mm

4. Pitch diameter,
p

5. Groove depth

= 2.065 x groove radius

h

= 0.374 x d

= 2.065 x 7.4mm

= 0.374 x 14.8

= 15.281mm

= 5.5352mm

6. Thickness
tx

7.
8.
9.
10.

= P/kp
= 15000/(210000000 x 0.0148)

Y= 520 mm
L5 = L6 = 20 mm
Number of layer = 3
Number of groove = 20

= 4.826mm
11. Drum grooved length,L3
L3

= (n - 1)P

= (20 – 1) 15.281

= 290.24 mm

12. Drum un-grooved length,L1=L2
L1 = L2

= 1/2 diameter of hook + radius of rope
Page 7

= 47.13 mm
13. Safe Factor
4.2

=6

SPECIFICATION OF DRUM

L1

L3 = (n – 1) P

L2

y

Ddrum

L6

L5

Figure 4: drum legend for revise


Page 8

Figure 5: groove overview


Page 9

4.3

DRAWING 2D

Figure 6: isometric drawing

Page 10

5.0

GEAR

5.1

DESIGN OVERVIEW

Figure 8: gear mechanism with legend


Page 11

5.2

CALCULATION FOR GEARS

For a single handle winch,
P=150N
a=350mm
D=280mm
W=15kN

Efficiency of gears and Drum,
n1 (G1, G2) =0.97
n2 (G3, G4) =0.97
n3 (drum) =0.96
n = n 1 x n2 x n 3
= 0.9033
Calculating for speed ratio, i,

(

)(
(


)(
)(

)
)

Page 12

Calculating suitable speed ratio for gear teeth,

*The highest speed ratio is taken to determine gear teeth.

Hence, calculating for the number of teeth of each gear,


Page 13

Calculating for new speed ratio value, i

By using new i value,

(

)(
( )(

)(
)(

)
)

 So these set of gear teeth is suitable for our winch design.


Page 14

Speed ratio:

Number of teeth (n) of gears:
Number of teeth of gear 1, Z1

=

11


=

70


=

11


=

77

*no. of teeth obtained by comparing with speed ratio
Efficiency of gear:
n1 [G1, G2] = 0.97
n2 [G3, G4] = 0.97
n3 [drum] = 0.96
n=n1 x n2 x n3

=0.9033

Pressure angle, Φ:
Φ = 20˚ (commonly used for designer)
Pitch diameter (D) of gears:
relationship:

Pitch diameter of gear 1, D1 = 66mm
Diametral Pitch (P):

Page 15

Example gear 1,= 11/ (66 x 0.0394)
For gear 1= 4.23 / inch
For gear 2= 4.23/ inch
Metric Module (m):

Applied for all gears, m = 6mm
Addendum (20o) (a):

Applied for all gears, a = 6.57mm
Dedendum (20o) (d):

Applied for all gears, b = 8.21mm
Outside Diameter (Do):
(

)

Gear 2 as example, Do = 432mm


Page 16

Torque, T
-applied to all gear

(

)(

)

P – obtained from the new i value.
a – length of handle.

Material:
Applied to all gear = AISI 1020 Steel
AISI 1020 Steel properties:

Figure 9: table properties of AISI 1020


Page 17

Face width (b):
By using Lewis Equation:
σ = (FtP)/(bY)
where,
σ - the allowable bending stress of the gear material
Y - the Lewis form factor which can be obtained from the standard charts with certain
pressure
angle, Φ
Ft - the tangential force applied to the gear tooth as show in the Figure 2.1 above
P - the diametral pitch
b – the face width
Rearranging the equation,
b=
Ft = F cos Φ

(F = 10 kN )

Ft = 15000 cos 20˚ = 14095.39 N. (apply to all gear)
P = 4.23 (apply to all gear)
σ = 124.2 MPa (For AISI 1020 steel)

Figure 10: Standard Lewis Form Factor chart

Page 18

Figure 11: lewis factor, Y table selection

For gear 1 & 3,
No. of teeth = 11, Y= 0.226

For gear 2,
No. of teeth = 70, Y= 0.429

For gear 4,
No. of teeth = 77, Y = 0.434
In order to have standard values of tooth width (b), we pick the smallest value of Lewis Form
factor Y = 0.226
b=
=
= 2.12mm
Take the value safety factor of 6 into considerations, the tooth width becomes 2.12 x 6 = 12.7mm

Page 19

5.3

TABLE OF OVERVIEW FOR GEARS
Detail

Gear 1 (Pinion)

Gear 2(Gear)

Gear 3 (Pinion)

Gear 4 (Gear)

Metric module (m)

6mm

6mm

6mm

6mm

Number of teeth (z)

11

70

11

77

Pressure angle (Φ)

20˚

20˚

20˚

20˚

Pitch Diameter (D)

66mm

420mm

66mm

462mm

Diametral Pitch (P)

4.23

4.23

4.23

4.23

Addendum (a)

6.57mm

6.57mm

6.57mm

6.57mm

Dedendum (d)

8.21mm

8.21mm

8.21mm

8.21mm

Efficiency (n)

0.9033

0.9033

0.9033

0.9033

Outside Diameter (Do)

78mm

432mm

78mm

474mm

AISI 1020 Steel

AISI 1020 Steel

Thickness

30mm

Material used
Speed ratio
Face width (b)

AISI 1020 Steel

AISI 1020 Steel

1/6.3267

1/7
12.7mm

Torque

Figure 12: Table showing the overview of the whole gear system

5.4

DRAWING 2D


Page 20

Figure 13: isometric view for gear one

Page 21


Figure 14: isometric view for gear two

Page 22


Figure 15: isometric view for gear three

Page 23


Figure 16: isometric view for gear four

Page 24


6.0

HANDLE

6.1

SPECIFICATIONS

Mechanical Properties
The mechanical properties of annealed AISI 1340 alloy steel are displayed in the following
table.
Properties

Metric

Imperial

Tensile strength

703 MPa

102000 psi

Yield strength

434 MPa

62900 psi

Bulk modulus (typical for steel)

140 GPa

20300 ksi

Shear modulus (typical for steel)

80.0 GPa

11600 ksi

190-210 GPa

27557-30458 ksi

0.27-0.30

0.27-0.30

Elongation at break (in 50 mm)

25.50%

25.50%

Reduction of area

57.30%

57.30%

46.0 J

33.9 ft-lb

Hardness, Brinell

207

207

Hardness, Knoop (converted from Brinell hardness)

207

207

Hardness, Rockwell B (converted from Brinell hardness)

93

93

Hardness, Rockwell C (converted from Brinell hardness)

15

15

217

217

50

50

Elastic modulus
Poisson's ratio

Izod Impact

Hardness, Vickers (converted from Brinell hardness)
Machinability (based on 100 machinability for AISI 1212 steel)

Figure 20: mechanical properties table for AISI 1340 alloy steel

Figure 21: element on handle

Page 25

6.2

Calculation

Tensile strength, Su

= 702 MPa

Yield strength , Sy

= 434 MPa

Design Factor = Safety Factor = 6

length, L = 300 mm

Maximum bending stress
μ = F.L = (150 N)(0.3m)
= 45 Nm
Compute the desire stress from;
σd = Sy / N
= 434 MPa / 6
= 72.333 MPa
Based on formula ; σ = μ / S

;

S=μ/σ
= 45Nm / 72.333MPa
= 6.2212 x 10 -7 m3
In realtionship between S, t and h is

; and h = 3t

S = (t.h2 ) / 6
= 9t3 / 6
= 1.5 t3
1.5 t3 = 6.2212 x 10 -7 m3
t = 7.4575 x 10 -3 m

therefore; h = 3t = 3(7.4575 x 10 -3 m) = 22.372 x 10 -3 m


Page 26

Figure 22: handle with labelling

ALL THE DATA;
t = 7.4575 x 10 -3 m

F = 0.30 m

h= 22.372 x 10 -3 m

E = 0.127 m


Page 27

6.3

2D DRAWING


Page 28

7.0

SHAFT

7.1

INTRODUCTION

In this project, we are focusing in winch to be able withstand load about 15kN. Other
than gears, bearings, drum, and other winch component. Shaft also plays as important factor to
increase efficiency and durability of winch.
In this report, we are discussed and find out the ideal diameter for shaft. For our winch we use 3
shafts to hold gears and transmit power

7.2

CALCULATION FOR SPEED 1

7.2.1 CALCULATION FOR SHAFT 1 AND DRAWING

Material selection

Referred back to the Stainless Steel 304, the material has possessed two properties which is
chemical and mechanical. In this scope, we more likely to know its mechanical properties which
are main reason due to the all the information that lead to the shaft requirement and can be
calculated based on formula that can be used. All the data for the chemical and mechanical
properties of the Stainless Steel 304can be summarize as seen in next page;


Page 29

Chemistry % by Weight
Element
Type 302

Type 304

Type 304L

Type 305

Fe

64.99-74%

Cr

17%

18%

18%

17%

Ni

8%

8%

8%

10.50%

Mn

2%

2%

2%

2%

N

0.10%

0.10%

0.10%

-

S

0.03%

0.03%

0.03%

0.03%

C

0.15%

0.08%

0.03%

0.12%

Si

0.75%

0.75%

0.75%

0.75%

P

0.045%

0.045%

0.045%

0.045%

Properties
Properties
Physical Properties
Density
Mechanical properties
Hardness, Rockwell B
Tensile Strength, Ultimate
Tensile Strength, Yield

Value

Comment

8.03 g/cc
82
621 Mpa (=90100psi)
290 Mpa (=42100psi)

0.2% YS

Elongation at Break

55%

in 2 inches

Modulus of Elasticity
Electrical properties
Electrical Resistivity
Magenetic permeability
Thermal properties
CTE, linear 20°C
CTE, linear 20°C
Heat Capacity
Thermal Conductivity
Processing properties
Melt temperature

193 Gpa
78 Gpa

tension
torsion

0.000116 ohm-cm
7.2e-005 ohm-cm
Max 1.02

659 °C

16.9 µm/m-°C
18.7 µm/m-°C
0.5 J/g-°C
16.2 W/m-K
21.4 W/m-K

0 to 100°C
to 649°C
0°C to 100°
100°C
500°C


H = 200 Oersteds, Annealed

1371- 1399 °C

Page 30

Endurance strength, Sn

= 0.5 Su
=0.5(621)
=310.50M Pa

Actual Endurance Strength, Sn'
Actual Endurance Strength, Sn’

= Sn (Cm) (Cst) (CR) (CS)

Cm = 1.0 (Since the material we used is steel)
Cst = 1.0 (Type of loading is bending)
CR = 0.81 (The shaft is designed to has reliability of 0.99)
CS = 0.75 (We assume it as we do not know the actual size yet)

=310.50 (1.0) x (1.0) x (0.81) x (0.75)
= 188.63MPa

Average force applied on handle is 150N, and the length of handle is 0.3m
Therefore, Torque, T = 150 x (0.3)
= 45N.m
Force on the Pinion 1, Tangential force, Wt =

=

= 1363.64 N


Page 31

Radial force, Wr
=Wt tan
=1363.64 tan 20 º
=496.32N

r1 is a sharp fillet, then a stress concentration factor of 2.5 should be used.
Safety Factor of 6 is used as the shaft is expected to withstand unusual shock or impact.
D1

=[

(

)

√(

)

( )

(

)

But the moment at point A is zero, thus

Page 32

D1

=[

( )

√( )

(

=20.17 mm = D4
D3

=[

(

)

√(

)

(

)

(this are due to the zero moment at both ends)
)

( )

(

)

Referring to the bending moment diagram above,
MBX = 42.42N m
MBY = 116.54 N m
Applying Pythagoras’s theorem, M = √
=√
= 124.02 N m

D3

=[

( )

√(

)

(

)

(

)

= 39.33 mm
D2 will be larger than D3 and D1 in order to provide a shoulder for the gear. The actual value of
D2 will be specified after the analysis has been completed.

Fillet designs for shaft 1
There are three fillets for shaft 1 which two of them are sharp and one of them is wellrounded. The value of Kt is 2.5 for sharp fillet and 1.5 for well-rounded fillet. For shaft 1, r1 and
r3 are shaft fillets while r2 is well-rounded fillet. By referring to figure below,


Page 33

Figure 1.2: Reference diagram used in determining the fillet radius

Stressconcentration

D/d

r/d

Fillet Radius

d (mm)

(mm)

factor, Kt
r1

2.5

2.27

0.015

22

0.33

r3

1.5

1.19

0.06

42

2.52

r4

2.5

1.91

0.015

22

0.33

Summary for shaft

1

Page 34

The calculated minimum required diameters for various parts of the shaft 1 are as follows:
D1

= 20.17 mm

D2

= 50.00 mm (D2 must be larger than D1 and D3 to provide shoulders for them)

D3

= 39.33 mm

D4

= 20.17 mm

The dimensions above are minimum diameters for various parts, to ease manufacturing process
and provide extra safety, rounded-off and larger diameters are used:
D1

= 22.00 mm

D2

= 50.00mm

D3

= 42.00mm

D4

= 22.00mm

Fillet Radius

Dimension (mm)

r1

0.33

r3

2.52

r4

0.33


Page 35


Gear 4 on shaft 2 receive the torque from pinion 1 on shaft 1.

Force on the Gear 4,
Tangential force, Wt(c)

=

=
= 1363.64 N


Page 37

Radial force, Wr (c)

=Wt tan
=1363.64 tan 20 º
=496.32 N

Force exerted by drum
WDrum = ρVg
= 8030 kg/m3 x 9.7839 x 10-3 m3 x 9.81 m/s2
= 770.72 N


Page 38

D1

=[

(

)

√(

)

( )

(

)

Moment is zero at both ends

D1

=[

( )

√( )

(

)

(

)

= 38.59 mm = D3

At point B, where there is a profile keyset at the gear 4.
Thus by calculation the theoretical Db is 45.64 mm
Diameter for 2 must be larger than D1 and D3. It also must exceed Db which is theoretical value
if gear is apart from the drum.

There are 2 fillets for shaft 2 which both are sharp. The value of Kt is 2.5 for sharp fillet.
For shaft 3, r1 and r3 are shaft fillets. By referring to figure below,


Page 39


Table 3.1: Fillet radius calculated based on the figure 3.1.
Stressconcentration

D/d

r/d

d (mm)

factor, Kt

Fillet Radius
(mm)

r1

2.5

1.5

0.015

40

0.6

r3

2.5

1.5

0.015

40

0.6


Page 40

Summary for shaft 2

D1

= 38.59mm

D2

= 60.00mm (D2 must be larger than D1 and D3 to provide shoulders for them)

D3

= 38.59mm

Also, on the left side of the Gear 4, there is where the drum placed. The drum exerts a force to
the shaft in vertical plane. Thus, D2 has to be increased again to ensure the force exerted from the
drum does not fail the shaft. All the dimensions above are then enlarged with certain ratio such
that the D2 has the same value as the hollow hole in the drum as what is in Drum Specification.
The values are the rounded-off in order the ease the manufacturing process.
D1

= 40.00mm

D2

= 60.00mm

D3

= 40.00mm
Fillet Radius

Dimension (mm)

r1

0.6

r3

0.6


Page 41


Torque was transmitted from gear 4 to the gear 3.


=

= 1363.63 N

Page 43


=Wt tan
=1363.63 tan 20 º
=496.32 N


=

= 214.29 N

=Wt tan
=214.29 tan 20 º
=77.994 N


Page 44

D1

=[

(

)

√(

)

( )

(

)

There has no moment at both ends

=[

( )

√( )

(

)

(

)

= 20.17 mm = D5
MBX = 44.64 N m
MBY = 122.69 N m

=√

)

(

)

= 130.56 N m

D3

=[

( )

√(

)

(

)

(

)

= 40.00 mm

McX = 16.698 N m
McY = 45.93 N m


Page 45

Applying Pythagoras’s theorem, Mc = √
)

= √(

(

)

= 48.87 N m

D4

=[

( )

√(

)

(

)

(

)

= 29.29 mm
There are four fillets for shaft 3 which two of them are sharp and two of them are wellrounded. The value of Kt is 2.5 for sharp fillet and 1.5 for well-rounded fillet. For shaft 2, r1 and
r4 are shaft fillets while r2 and r3 are well-rounded fillets. By referring to figure below,



Page 46

Table 2.1: Fillet radius calculated based on the Figure 2.1.
Stressconcentration

D/d

r/d

d (mm)

factor, Kt

Fillet Radius
(mm)

r1

2.5

2.5

0.015

22

0.33

r3

1.5

1.25

0.07

44

3.08

r4

1.5

1.33

0.08

33

2.64

r5

2.5

1.5

0.015

22

0.33

Summary for shaft 3



Page 47

D1

= 20.17mm

D2


D3

= 40.00 mm

D4

= 29.29mm

D5

= 20.17mm

D1

= 22.00 mm

D2

= 55.00mm

D3

= 44.00mm

D4

= 33.00mm

D5

= 22.00mm
Table 2.2: Fillet radii used for shaft 2
Fillet Radius

Dimension (mm)

r1

0.33

r3

3.08

r4

2.64

r5

0.33


Page 48

7.3

KEY AND KEYSEATS CALCULATIONS
In order to allow the gear transmit its rotational power to the shaft effectively and without

any slipping, keys are used for all the pinions and gears in our design. In our design, profile
rectangular keys are used as they are most common used keys in industry as well as due to a
lower cost required. Profile rectangular keys can ensure that the shaft and gear interlock each
other in all direction, which is better to prevent the gear from moving in any direction.


Page 50

Shaft 1, for gear 1

H = 9mm

W = 14mm

Shaft 2, for gear 4

H = 11mm

W = 18mm

Shaft 3, for gear 3and 2

H = 9mm

W = 14mm

Gear 3

H = 8mm

W = 10mm

Gear 2


Page 52

Key for gear 1
Bearing and Shear Stress Analysis
Since for gear 1, the diameter of shaft where the gear mounted on it is 42.00mm. From Appendix
in page 23, the standard width and height of the key are 14.00mm and 9.00mm respectively.
Width of the key, W = 14.00mm
Diameter of the shaft, D = 42.00mm
The key is expected to withstand sudden shock or impact, thus the safety factor used is 6.
From calculation for shaft 1, the torque transmitted is 45.0 N m
Tensile strength of shaft 1, Su = 621 M Pa.
The minimum required key length can be calculated using the formula as below:

Ls

=
= 2.96 mm

Key for gear 2
For gear 2, the diameter of shaft where the gear 2 mounted on it is 33.00mm. From Appendix in
page 23, the standard width and height of the key are 10.00mm and 8.00mm respectively. Width
of the key, W = 10.00mm
From calculation for shaft 1, the torque transmitted is 45.0N m


Page 53


Ls

=
= 5.27 mm

Key for gear 3
page 23, the standard width and height of the key are 14.00mm and 9.00mm respectively. Width
From calculation for shaft 3, the torque transmitted is 45 N m

Ls

=
= 2.82 mm

Key for gear 4
page 23, the standard width and height of the key are 18.00mm and 11 mm respectively. Width

Page 54

From calculation for shaft 2, the torque transmitted is 315 N m


Ls

=
= 11.27 mm

The values of required length above are the minimum lengths required for every single key. The
values are enlarged to ease the manufacturing process and provide extra safety.

Summary
Table 3.1: Dimensions of keys used
Torque

Required

Transmitted

length of key

(N m)

(mm)

14.00

45.0

2.96

30.00

33.00

10.00

45.0

5.27

30.00

Gear 3

44.00

14.00

45.0

2.82

30.00

Gear 4

60.00

18.00

315.0

11.27

30.00

Diameter of

Width of key

Shaft (mm)

(mm)

Gear 1

42.00

Gear 2


Actual length
of key (mm)

Page 55

7.4

CALCULATION FOR SPEED 2


Material selection
Referred back to the Stainless Steel 304, the material has possessed two properties which is
chemical and mechanical. In this scope, we more likely to know its mechanical properties which
are main reason due to the all the information that lead to the shaft requirement and can be
calculated based on formula that can be used. All the data for the chemical and mechanical
properties of the Stainless Steel 304can be summarize as seen in next page;

Element
Type 302

Type 304

Type 304L

Type 305

Fe

64.99-74%

Cr

17%

18%

18%

17%

Ni

8%

8%

8%

10.50%

Mn

2%

2%

2%

2%

N

0.10%

0.10%

0.10%

-


Page 56

S

0.03%

0.03%

0.03%

0.03%

C

0.15%

0.08%

0.03%

0.12%

Si

0.75%

0.75%

0.75%

0.75%

P

0.045%

0.045%

0.045%

0.045%

Properties
Properties
Physical Properties
Density

Value

Comment

8.03 g/cc
82
621 Mpa (=90100psi)
290 Mpa (=42100psi)

0.2% YS

Elongation at Break

55%

in 2 inches

Thermal properties
CTE, linear 20°C
CTE, linear 20°C
Heat Capacity
Melt temperature

193 Gpa
78 Gpa

tension
torsion

0.000116 ohm-cm
7.2e-005 ohm-cm
Max 1.02

659 °C

16.9 µm/m-°C
18.7 µm/m-°C
0.5 J/g-°C
16.2 W/m-K
21.4 W/m-K

0 to 100°C
to 649°C
0°C to 100°
100°C
500°C


1371- 1399 °C


= 0.5 Su
=0.5(621)
=310.50M Pa


Page 57

Actual Endurance Strength, Sn'

= Sn (Cm) (Cst) (CR) (CS)

Cm = 1.0 (Since the material we used is steel)
Cst = 1.0 (Type of loading is bending)
CR = 0.81 (The shaft is designed to has reliability of 0.99)
CS = 0.75 (We assume it as we do not know the actual size yet)

=310.50 (1.0) x (1.0) x (0.81) x (0.75)
= 188.63MPa

Average force applied on handle is 150N, and the length of handle is 0.3m
Therefore, Torque, T = 150 x (0.3)
= 45N.m
Force on the Pinion 1, Tangential force, Wt =

=
= 1363.64 N

Radial force, Wr

=Wt tan
=1363.64 tan 20 º
=496.32N


Page 58

r1 is a sharp fillet, then a stress concentration factor of 2.5 should be used.
Safety Factor of 6 is used as the shaft is expected to withstand unusual shock or impact.
D1

=[

(

)

√(

)

( )

(

)

(

)

But the moment at point A is zero, thus

D1

=[

( )

√( )

(

=20.17 mm = D4
D3

=[

(

)

√(


)

(this are due to the zero moment at both ends)
)

( )

(

)

Page 59

MBX = 16.29N m
MBY = 44.77 N m
=√
= 47.64 N m

D3

=[

( )

√(

)

(

)

(

)

= 29.07 mm
D2 will be larger than D3 and D1 in order to provide a shoulder for the gear. The actual value of
D2 will be specified after the analysis has been completed.

There are three fillets for shaft 1 which two of them are sharp and one of them is wellrounded. The value of Kt is 2.5 for sharp fillet and 1.5 for well-rounded fillet. For shaft 1, r1 and
r3 are shaft fillets while r2 is well-rounded fillet. By referring to figure below,


Page 60

Stressconcentration

D/d

r/d

d (mm)

factor, Kt

Fillet Radius
(mm)

r1

2.5

2.27

0.015

22

0.33

r3

1.5

1.21

0.06

33

1.98

r4

2.5

1.91

0.015

22

0.33

Summary for shaft 1

D1

= 20.17 mm

D2


D3

= 29.07 mm

D4

= 20.17 mm


Page 61

The dimensions stated previous page are minimum diameters for various parts, to ease
manufacturing process and provide extra safety, rounded-off and larger diameters are used:
D1

= 22.00 mm

D2

= 40.00mm

D3

= 33.00mm

D4

= 22.00mm

Fillet Radius

Dimension (mm)

r1

0.33

r3

1.98

r4

0.33


Page 62


Gear 4 on shaft 2 receive the torque from pinion 1 on shaft 1.


=

=
= 8677.58 N


Page 64


=Wt tan
=8677.58 tan 20 º
=3158.38 N

Force exerted by drum
WDrum = ρVg
= 8030 kg/m3 x 9.7839 x 10-3 m3 x 9.81 m/s2
= 770.72 N

D1

=[

(

)

√(


)

( )

(

)

Page 65

Moment is zero at both ends

D1

=[

( )

√( )

(

)

(

)

= 71.52 mm = D3

At point B, where there is a profile keyset at the gear 4.
Thus by calculation the theoretical Db is 73.62 mm
Diameter for 2 must be larger than D1 and D3. It also must exceed Db which is theoretical value
if gear is apart from the drum.

There are 2 fillets for shaft 2 which both are sharp. The value of Kt is 2.5 for sharp fillet.
For shaft 3, r1 and r3 are shaft fillets. By referring to figure below,



Page 66

Table 3.1: Fillet radius calculated based on the figure 3.1.
Stressconcentration

D/d

r/d

d (mm)

Fillet Radius

factor, Kt

(mm)

r1

2.5

1.13

0.01

75

0.75

r3

2.5

1.13

0.01

75

0.75

Summary for shaft 2

D1

= 71.52mm

D2

= 85.00mm (D2 must be larger than D1 and D3 to provide shoulders for them)

D3

= 71.52mm


Page 67

Also, on the left side of the Gear 4, there is where the drum placed. The drum exerts a force to
the shaft in vertical plane. Thus, D2 has to be increased again to ensure the force exerted from the
drum does not fail the shaft. All the dimensions above are then enlarged with certain ratio such
that the D2 has the same value as the hollow hole in the drum as what is in Drum Specification.
The values are the rounded-off in order the ease the manufacturing process.
D1

= 75.00mm

D2

= 85.00mm

D3

= 75.00mm
Fillet Radius

Dimension (mm)

r1

0.75

r3

0.75


Page 68


Torque was transmit from gear 1 to the gear 2.


=

= 1363.63 N

Page 70


=Wt tan
=1363.63 tan 20 º
=496.32 N


=

= 8677.57 N

=Wt tan
=8677.57 tan 20 º
=3158.38 N


Page 71

D1

=[

(

)

√(

)

( )

(

)

There has no moment at both ends

=[

( )

√( )

(

)

(

)

= 37.39 mm = D5
Referring to the bending moment diagram above, for at point gear 3
MBX = 284.52N m
MBY = 780.52 N m

= √(

)

(

)

= 830.76 N m

D3

=[

( )

√(

)

(

)

= 74.09 mm

Referring to the bending moment diagram above, at point gear 2
McX = 106.78 N m
McY = 291.99 N m


Page 72

Applying Pythagoras’s theorem, Mc = √
= √(

)

(

)

= 310.90 N m

D4

=[

( )

√(

)

(

)

= 54.26 mm
There are four fillets for shaft 3 which two of them are sharp and two of them are wellrounded. The value of Kt is 2.5 for sharp fillet and 1.5 for well-rounded fillet. For shaft 2, r1 and
r4 are shaft fillets while r2 and r3 are well-rounded fillets. By referring to figure below,



Page 73

Table 2.1: Fillet radius calculated based on the Figure 2.1.
Stressconcentration

D/d

r/d

d (mm)

factor, Kt

Fillet Radius
(mm)

r1

2.5

2.0

0.015

40

0.6

r3

1.5

1.067

0.025

75

1.875

r4

1.5

1.36

0.075

55

4.125

r5

2.5

1.375

0.015

40

0.6

Summary for shaft 3


Page 74


D1

= 37.39mm

D2


D3

= 74.09 mm

D4

= 54.26mm

D5

= 37.39mm

D1

= 40.00 mm

D2

= 80.00mm

D3

= 75.00mm

D4

= 55.00mm

D5

= 40.00 mm
Table 2.2: Fillet radii used for shaft 2
Fillet Radius

Dimension (mm)

r1

0.6

r3

1.875

r4

4.125

r5

0.6


Page 75

7.5

KEY AND KEYSEATS CALCULATIONS
In order to allow the gear transmit its rotational power to the shaft effectively and without

any slipping, keys are used for all the pinions and gears in our design. In our design, profile
rectangular keys are used as they are most common used keys in industry as well as due to a
lower cost required. Profile rectangular keys can ensure that the shaft and gear interlock each
other in all direction, which is better to prevent the gear from moving in any direction.


Page 77

Shaft 1, for gear 1

H = 8mm

W = 10mm

Shaft 2, for gear 4

H = 14mm

W = 25mm

Shaft 3, for gear 3and 2

H = 14mm

W = 22mm

Gear 3

H = 10mm

W = 16mm

Gear 2


Page 79

Key for gear 1
Since for gear 1, the diameter of shaft where the pinion 1 mounted on it is 33.00mm. From
Appendix in page 23, the standard width and height of the key are 10.00mm and 8.00mm
respectively.

Ls

=
= 5.27 mm

Key for gear 2
Since for gear 2, the diameter of shaft where the gear 2 mounted on it is 55.00mm. From
Appendix in page 23, the standard width and height of the key are 16.00mm and 10.00mm
respectively. Width of the key, W = 16.00mm


Page 80


Ls

=
= 12.58 mm

Key for gear 3
page 23, the standard width and height of the key are 22.00mm and 14.00mm respectively.

Ls

=
= 6.71 mm

Key for gear 4
page 23, the standard width and height of the key are 25.00mm and 14 mm respectively. Width

Page 81



Ls

=
= 36.46 mm

The values of required length above are the minimum lengths required for every single key. The
values are enlarged to ease the manufacturing process and provide extra safety.

Summary
Table 3.1: Dimensions of keys used
Torque

Required

Transmitted

length of key

(N m)

(mm)

10.00

45.0

5.27

30.00

55.00

16.00

286.36

12.58

30.00

Gear 3

75.00

22.00

286.36

6.71

30.00

Gear 4

85.00

25.00

2004.52

36.46

30.00

Diameter of

Width of key

Shaft (mm)

(mm)

Gear 1

33.00

Gear 2


Actual length
of key (mm)

Page 82

7.6

CONCLUSION OF SHAFT

From the calculation for both speed as we calculated as shown in previous page, the shaft
diameter calculated give the difference between both speeds. Therefore, we are considering the
shaft with bigger diameter for our selection due to safety reason.
In comparison, shaft 1 for speed one is bigger than diameter of shaft for speed 2. Thus we
are select diameter in calculation for speed 1. For shaft 2, diameter calculated for speed 1 smaller
than diameter calculated for speed 2. The diameter calculated for speed 2 is selected. For the
third shaft, the diameter calculated for speed 1 also smaller than diameter for shaft calculated
from speed 2. Thus we select diameter calculated for the speed 2.
The final selection of the shaft is listed as below for our design of winch.
SHAFT NO.
1

Diameter 1
(mm)
22.00

Diameter 2
(mm)
50.00

Diameter 3
(mm)
42.00

Diameter 4
(mm)
22.00

Diameter
5(mm)
-

2

75.00

85.00

75.00

-

-

3

40.00

80.00

75.00

55.00

40.00


Page 83

8.0

LUBRICATION

8.1

INTRODUCTION OF LUBRICATION
Lubrication is simply the use of a material to improve the smoothness if movement of

one surface over another; the material which is used in this way is called a lubricant. Lubricants
are usually liquids or semi-liquids, but may be solids or gases or any combination of solids,
liquids, and gases. The smoothness of movement is improved by reducing friction. This is not,
however, always the case, and there may be situations in which it is more important to maintain
steady friction than to obtain the lowest possible friction. In addition to reducing or controlling
friction, lubricants are usually expected to reduce wear and often to prevent overheating and
corrosion.
TYPES OF LUBRICANT
Lubricants are usually divided into four basic classes.
No.
1

Type
Oils

Description
A general term used to cover all liquid lubricants, whether they are
mineral oils, natural oils, synthetics, emulsions, or even process fluids.

2

Greases

Technically these are oils, which contain a thickening agent to make
them semi-solid. It is convenient, however, to include the anti-seize
pastes and the semi-fluid greases under the same heading.

3

Dry lubricant

These include any lubricants, which are used in solid form, and may be
bulky solids, paint-like coatings, or loose powders.

4

Gases

The gas usually used in gas bearings is air, but any gas can be used
which will not attack the bearings, or itself decompose.
Table 1: type of lubricant in industry

OIL LUBRICATION

Oils are used in journal bearings when cooling is required or contaminants or debris need to be

Page 84

flushed away from the bearing. High-speed journal bearings are always lubricated with oil rather
than grease. Oil is supplied to the bearing by either a pressurized oil pump system, an oil ring or
collar or a wick. Grooves in the bearing shell are used to distribute the oil throughout the
bearings’ surfaces.
The viscosity grade required is dependent upon bearing RPM, oil temperature and load. The
bearing speed is often measured strictly by the revolutions per minute of the shaft, with no
consideration of the surface speed of the shaft.
Bearing Speed
(rpm)

Bearing / Oil Temperature (°C)
0 to 50

60

75

90

-

68

100 to 150

-

~1,800

32

32 to 46

68 to 100

100

~3,600

32

32

46 to 68

68 to 100

~10,000

32

32

32

32 to 46

300 to 1,500

Table 2: Journal Bearing ISO Viscosity Grade Selection


Page 85

9.0

BEARING

9.1

INTRODUCTION OF BEARING
Bearing is machine element that constrains relative motion and reduces friction

between moving parts to only the desired motion. The design of the bearing may, for example,
provide for free linear movement of the moving part or for free rotation around a fixed axis; or, it
may prevent a motion by controlling the vectors of normal forces that bear on the moving parts.
Many bearings also facilitate the desired motion as much as possible, such as by minimizing
friction. Bearings are classified broadly according to the type of operation, the motions allowed,
or to the directions of the loads (forces) applied to the parts.
TYPES OF BEARING
Type
Plain bearing

Description
Rubbing surfaces, usually with lubricant; some bearings use pumped
lubrication and behave similarly to fluid bearings.

Rolling element

Ball or rollers are used to prevent or minimize rubbing

bearing
Jewel bearing

Off-center bearing rolls in seating

Fluid bearing

Fluid is forced between two faces and held in by edge seal

Magnetic bearing

Faces of bearing are kept separate by magnets

Flexure bearing

Material flexes to give and constrain movement
Table 3: type of bearing in industry

For our design we are focusing in using plain bearing which is journal type, straight
journal bearing. This type of journal bearing operates in the boundary regime (metal-to-metal
contact) only during the startup and shutdown of the equipment when the rotational speed of the
shaft (journal) is insufficient to create an oil film. As we known, this type of journal is selflubricate.


Page 86

It is important to understand that the rotating shaft is not centered in the bearing shell during
normal operation. This offset distance is referred to as the eccentricity of the bearing and creates
a unique location for the minimum oil film thickness, as shown in next page.

Figure 1: operation of lubricant and journal bearing


Page 87

9.2

DESIGN OF BEARING
To get optimal performance from the bearing, designer must pay special attention to

design as regards the fit between the inner ring, outer ring, and the housing. This fit and design
must be suitable for the specific equipment involved in the application. For example, too loose a
fit could result in a corroded or scored bearing bore and shaft; while too tight a fit could result in
unnecessarily large mounting and dismounting forces and too great a reduction in internal
bearing clearance.
All rolling bearing manufacturers make bearings to standardized tolerances set forth by
the Anti-Friction Bearing Manufacturers Association (AFBMA) and the International Standards
Organization (ISO). The proper fits can only be obtained by selecting the proper tolerances for
the shaft. Each tolerance is designated by a letter and a numeral. The small letter is for shaft fits
and the capital letter is for housing bores, and together they locate the tolerance zone in relation
to the nominal dimensions. The numeral designation gives the magnitude of the tolerance zone.
Shaft and Housing
Finish precision of bearing shaft and housing:
After bearing installation, degradation will occur in the roundness of the bearings if the
design and fit of the shaft and housing are not appropriate. The precision and surface roughness
of the shaft and housing must be at satisfactory levels to protect the bearings and optimize their
performance.
Fillet radii of corners of shaft and housing:
The design of the side faces of bearing shafts and housings, or the area contact the
bearing side face, causes them to come to the shaft center and fit surface at right angles. The
fillet radii of the shaft and housing corners have a maximum permissible radius (ras max) is
smaller than minimum permissible chamfer dimensions of the bearings.


Page 88

Height of shoulder:
The design of the shaft and housing shoulders must include a height that is taller than
minimum permissible chamfer dimensions. The shoulders also must be designed to contact the
side faces of the inner and outer rings.
To design shoulders properly, ensure that the minimum height of shoulders is four times
taller than minimum permissible chamfer dimensions. The table and figure on this page provide
additional insight into the dimension specifics of this bearing design requirement.
Referred from JIS B 1566

Figure 3: shaft housing and chamfer


Page 89

Bearing Material
Bearing materials constitute an import part of any bearing. Their significance is at the start of the
hydro-dynamic lubrication when metal to metal contact occurs or during mixed and boundary
lubrication period.
Desirable properties of good bearing material are;1. Conformability (low elastic modulus) and deformability (plastic flow) to relieve local
high pressures caused by misalignment and shaft deflection.
2. Embeddability or indentation softness, to permit small foreign particles to become
safely embedded in the material, thus protecting the journal against wear.
3. Low shear strength for easy smoothing of surface asperities.
4. Adequate compressive strength and fatigue strength for supporting the load and for
enduring the cyclic loading as with engine bearings under all operating conditions.
5. Should have good thermal conductivity to dissipate the frictional heat and coefficient
of thermal expansion similar to the journal and housing material.
6. It should be compatible with journal material to resist scoring, welding and seizing.
7. Should have good corrosion resistance against the lubricant and engine combustion
products.
Babbits are the most commonly used bearing materials. Babbitts have excellent
conformability and embeddability, but have relatively low compressive and fatigue
strength.Other materials such as tin bronze, leaded bronze, copper lead alloy, aluminium bronze,
aluminium alloys and cast iron are also used in many applications. Widely used bearing material
compositions are given below:
a.Tin-base babbitts with 89% Sn, 8% Pb and 3% Cu,
b. Lead- base babbitts with 75% Pb, 15% Sb and 10% Sn,
c. Copper alloys such as Cu- 10% to 15% Pb.


Page 90

Bimetal and trimetal bearings are used in engine application to reduce the size of the
bearing and obtain good compatibility and more load capacity. The bearings can be of solid
bushings or lined bushings. Sometimes two piece with or without flanges are also used. These
are shown in Fig.4.1. The inner surfaces of the bearings are grooved to facilitate the supply of
lubricant to the surface of the journal. Various groove pattern used in industry are shown in Fig.
4.2.

a) Solid bushing

b) Fanged

c) Lined bushing

d) Straight
Figure 4.1: various types of bush bearing


Page 91

Figure 4.2: developed views of typical groove patterns

Bearing Life
Generally, a rolling bearing cannot rotate forever. Unless operating conditions are ideal
and the fatigue load limit is not reached, sooner or later material fatigue will occur. The period
until the first sign of fatigue appears is a function of the number of revolutions performed by the
bearing and the magnitude of the load. Fatigue is the result of shear stresses cyclically appearing
immediately below the load carrying surface. After a time these stresses cause cracks which
gradually extend up to the surface. As the rolling elements pass over the cracks fragments of
material break away and this is known as flaking or spalling.
The flaking progressively increases in extent and eventually makes the bearing unserviceable (fig
5.1, fig 5.2, fig 5.3, fig 5.4)

fig 5.1


fig 5.2

fig 5.3

fig 5.4

Page 92

The life of a rolling bearing is defined as the number of revolutions the bearing can perform
before incipient flaking occurs. This does not mean to say that the bearing cannot be used after
then. Flaking is a relatively long, drawn-out process and makes its presence known by increasing
noise and vibration levels in the bearing. Therefore, as a rule, there is plenty of time to prepare
for a change of bearing.


Page 93

9.3

BEARING SELECTION


Page 94

9.4

2D DRAWING
The bearing was selected by refereed to the diameter of shaft. The diameter that was used

the bearing was the diameter of the ends of every shaft, which is the required internal diameter
for the bearing. The picture only show one drawing due to the same design, but only the inner
and outer diameter is different for every shaft.


Page 95

10.0

FRAME

10.1

INTRODUCTION
The Stainless Steel 304, the material has possessed two properties which is chemical and

mechanical. In this scope, we more likely to know its mechanical properties which are main
reason due to the all the information that lead to the shaft requirement and can be calculated
based on formula that can be used. All the data for the chemical and mechanical properties of the
Stainless Steel 304can be summarize as seen below;
The selection was influenced by the material shaft, this are due to the safety issues, which
is the component will hold the same strain and the limit also same, thus the safety issue does be
easier to counter.

Element
Type 302

Type 304

Type 304L

Type 305

Fe

64.99-74%

Cr

17%

18%

18%

17%

Ni

8%

8%

8%

10.50%

Mn

2%

2%

2%

2%

N

0.10%

0.10%

0.10%

-

S

0.03%

0.03%

0.03%

0.03%

C

0.15%

0.08%

0.03%

0.12%

Si

0.75%

0.75%

0.75%

0.75%

P

0.045%

0.045%

0.045%

0.045%


Page 96

Properties
Properties
Physical Properties
Density

Value

Comment

8.03 g/cc
82
621 Mpa (=90100psi)
290 Mpa (=42100psi)

0.2% YS

Elongation at Break

55%

in 2 inches

Thermal properties
CTE, linear 20°C
CTE, linear 20°C
Heat Capacity
Melt temperature

193 Gpa
78 Gpa

tension
torsion

0.000116 ohm-cm
7.2e-005 ohm-cm
Max 1.02

659 °C

16.9 µm/m-°C
18.7 µm/m-°C
0.5 J/g-°C
16.2 W/m-K
21.4 W/m-K

0 to 100°C
to 649°C
0°C to 100°
100°C
500°C



1371- 1399 °C

Page 97

10.2

2D DRAWING


Page 98

11.0

RATCHET AND PAWL

11.1

INTRODUCTION
Ratchets and pawls are mechanical assemblies allowing a shaft to rotate in one direction

and not the other. Ratchets are commonly known as ratchet wheels, because they consist of a
rotational gear (or rack) with angled teeth. Pawls are thin pieces of metal which rest against
ratchets to restrict the ratchet's motion. When the ratchet is rotated in one direction, the pawl is
raised and moves smoothly between the ratchet's teeth. When the ratchet's rotation stops, the
pawl drops between the ratchet teeth and makes clicking noises.
Ratchets can operate quietly and smoothly or be harsh and jarring, depending on the installation
of the ratchet teeth and pawl. Common applications for ratchets and pawls include use in
spanners, turnstiles, winders and jacks.

Figure 1: Ratchet and Pawl
In our winch design, ratchet will attach to Gear B (refer Figure 2). The force manually
applied on handle will be transmitted to gear A, which is 150N. The efficiency of Gear A and B
is 97%. So the force generated on Gear B as well as the ratchet is 150N x 0.97 = 145.5N.


Page 99

A

D

B

C

Figure 2: Gears Design
Referring to website catalogue KHK STOCK GEARS, ratchet (SRTB3-40) is selected.
The pitch for the ratchet is 9.42. The selection criteria are based on calculation below:
Allowable Torque = Allowable Force x Distance
Allowable Force = Allowable Torque / Distance = 157.52/0.3 = 525.07
Safety Factor = 525.07 / 145.2 = 3.61
Pawl model SRT-2C is selected as it works together with ratchet SRTB3-40. As refer to
the catalogue, the material for both ratchet and pawl is carbon steel S45C. The material
properties for S45C are shown below:
1. S45C Mechanical properties:
a) Density: 7700 – 8030 kg/m³
b) Young’s Modulus: 190-210 GPa
c) Tensile strength: 569MPa (Standard), 686MPa (Quenching, Tempering)
d) Yield Strength: 343MPa (Standard), 490MPa (Quenching, Tempering)
e) Poisson’s ratio: 0.27-0.30


Page 100

11.2

CATALOG

Catalog No.
Angle of
jaw
or teeth
Material
Heat
treatment
Surface
hardness
Surface
treatment
Surface
finish
Datum
reference
surface
for teeth
cutting


Specifications
SRT-C

SRT/SRTB

60

60

S45C (Carbon steel) **
Induction hardened
teeth
HRC48 53

S45C (Carbon steel)
Induction hardened
teeth
HRC48 53

Black oxide

Black oxide

Cut

Cut

Bore

Bore

Page 101

Pawls
Catalog No.
SRT2/3-C

J
5

K
8

L
38

M
30

N
6

Shape Weight(kgf)
T5
0.02

8

10

49

39 12

T5

0.05

10 12.5 67.5 55 15

T5

0.12

12

15

80

65 20

T5

0.22

13

18

98

80 25

T5

0.3

[DXF | DWG | 3D CAD]

SRT1-C

SRT2-C

SRT3-C

SRT4-C

Ratchets

Catalog
No.

Pit
ch

P

SRT1-50
[DXF | DW
G | 3D
CAD]

SRT1-60
[DXF | DW
G | 3D
CAD]

SRT1-80
[DXF | DW
G | 3D
CAD]

SRT1-90
[DXF | DW
G | 3D
CAD]

SRT1100

3.
14

No
.
of
te
et
h
z

B
or
e

Outsi
de
dia.

Face
widt
h

Dep
th
of
teet
h

Mou
ntin
g
dista
nce

Center
distanc
e

A

C

F

G

H

I

50

1
2

50

12

1.6

23.4

45.5

60

28.4

80

60
80
90
10
0

We
igh
t
(kg
f)

Catal
og
No.

0.1
6

SRT150

0.2
4

SRT160

0.4
4

SRT180

1
5

0.5
6

SRT190

1
5

0.7

SRT1100

1
5
1
5

Sh
ap
e

Allowa
ble
torque
(kgf-m)

Allow
able
torqu
e (Nm)

Bendin
g
strengt
h
1.5

Bendi
ng
streng
th
14.69

48.2

1.99

19.5

38.4

54.7

3

29.37

90

43.4

58.3

3.52

34.47

100

48.4

62.2

4.02

39.4

T4

[DXF | DW
G | 3D
CAD]


Page 102

SRT2-30
[DXF | DW
G | 3D
CAD]

6.
28

SRT2-40

30
40
50

[DXF | DW
G | 3D
CAD]

60

SRT2-50
[DXF | DW
G | 3D
CAD]
[DXF | DW
G | 3D
CAD]
[DXF | DW
G | 3D
CAD]

9.
42

SRT3-40

30
40
50

[DXF | DW
G | 3D
CAD]
[DXF | DW
G | 3D
CAD]
[DXF | DW
G | 3D
CAD]

SRT4-40
[DXF | DW
G | 3D
CAD]

SRT4-50
[DXF | DW
G | 3D
CAD]

1
5

15

3.1

26.9

61.2

80

36.9

100
120

T4

2.96

29.03

66.2

5.02

49.22

46.9

72.3

7.22

70.82

56.9

79.1

9.61

94.28

0.2
8

SRT230

0.5
3

SRT240

0.8
6

SRT250

1.2
4

1
5
2
0

90

20

5

40

76.3

120

55

150

70

T4

SRT260

0.8
6

SRT330

1.5
8

SRT340

2.5
4

SRT350

225.9
2

1.9
1

SRT430

385.1
4

3.5
4

SRT440

559.4

5.6
8

SRT450

9.44

92.56

85.1

16.06

95.5

23.31

157.5
2
228.6
2

2
0

SRT3-50

SRT4-30

1
5

60

1
5

SRT2-60

SRT3-30

1
5

12
.5
6

30
40
50

2
0
2
0
2
0


120

25

7.4

52.6

95.7

T4

23.04

160

72.6

108

39.27

200

92.6

122.4

57.04

Page 103

Ratchets with Hubs

Catalo
g No.

[DXF | D
WG | 3D
CAD]

N
o.
of
te
et
h

B
o
r
e

Hu
b
dia
.

Out
side
dia.

Fac
e
wid
th

Hu
b
wid
th

Tot
al
leng
th

De
pth
of
tee
th

Mo
unt
ing
dist
anc
e

Cent
er
dista
nce

P

SRTB2
/3-50

Pi
tc
h

z

A

B

C

D

E

F

G

H

I

2.
0
9

5
0

1
0

25

33.3

6

10

16

1

33.8

30

40

15.
5

6
0

1
0

8
0

1
2

9
0

1
2

1
0
0

1
2

SRTB2
/3-60
[DXF | D
WG | 3D
CAD]

SRTB2
/3-80
[DXF | D
WG | 3D
CAD]

SRTB2
/3-90

35

19

53.3

40

60

25.
5

40

66.6

29

35.5

S
h
a
p
e

T
9

Allo
wabl
e
torq
ue
(kgfm)
Bend
ing
stren
gth
0.31
0.42

Allo W Cata
wabl ei
log
e
gh No.
torq t
ue
(k
(N- gf)
m)
Ben
ding
stre
ngth
3.07 0. SRT
06 B2/3
4.1 7
-50

39.4

0.61

6

41.7

0.73

7.11

43.9

0.84

8.24

32

0.
1
0.
16
0.
21
0.
24

[DXF | D
WG | 3D
CAD]

SRT
B2/3
-60
SRT
B2/3
-80
SRT
B2/3
-90
SRT
B2/3
-100

SRTB2
/3100
[DXF | D
WG | 3D
CAD]

SRTB1
-50
[DXF | D
WG | 3D
CAD]

SRTB1
-60

3.
1
4

5
0

1
2

6
0

1
5

35

50

40

60

50

80

50


90

12

12

24

1.6

23.
4
28.
4

45.5
48.2
54.7
58.3

T
9

1.5
1.99
3
3.52

14.6
9

0.
24

19.5

0.
34

29.3

SRT
B150
SRT
B1-

Page 104

[DXF | D
WG | 3D
CAD]

8
0
9
0

[DXF | D
WG | 3D
CAD]

SRTB1
-90
[DXF | D
WG | 3D
CAD]

1
5

43.
4

1
0
0

SRTB1
-80

1
5

1
5

48.
4

50

38.
4

100

62.2

4.02

7
34.4
7
39.4

0.
61
0.
73
0.
87

[DXF | D
WG | 3D
CAD]

[DXF | D
WG | 3D
CAD]

6.
2
8

SRTB2
-50

1
5

4
0

1
5
1
5

6
0

[DXF | D
WG | 3D
CAD]

3
0

5
0

SRTB2
-40

50

60

60

80

1
5

60

14

29

3.1

26.
9
36.
9

100

65

15

120

46.
9

61.2

2.96
5.02

72.3

7.22

79.1

9.61

56.
9

29.0
3

0.
48

49.2
2

0.
82

70.8
2

1.
14

94.2
8

66.2

T
9

1.
59

[DXF | D
WG | 3D
CAD]

[DXF | D
WG | 3D
CAD]

[DXF | D
WG | 3D
CAD]

9.
4
2

[DXF | D
WG | 3D
CAD]

3
0

1
5

4
0

2
0

5
0

SRTB3
-40

75

90

80
85

20

16

36

5

40

76.3

120

55

85.1

150

70

T
9

95.5

2
0

9.44
16.0
6
23.3
1

92.5
6

1.
4

157.
52

2.
17

228.
62

3.
21

[DXF | D
WG | 3D
CAD]

[DXF | D
WG | 3D
CAD]

SRTB4
-40
[DXF | D

SRT
B230
SRT
B240
SRT
B250

SRT
B330
SRT
B340
SRT
B350

SRTB3
-50

SRTB4
-30

SRT
B190

SRT
B260

SRTB2
-60

SRTB3
-30

SRT
B180

SRT
B1100

SRTB1
-100

SRTB2
-30

60

1
2.
5
6

3
0

2
0

4
0

2
0

5

90

120

90

160

10
0

200

2


25

18

43

7.4

52.
6
72.
6
92.

95.7
108
122.4

T
9

23.0
4

225.
92

2.
76

39.2
7

385.
14

4.
4

57.0

559.

6.

SRT
B430
SRT
B440

Page 105

WG | 3D
CAD]

0

0

SRTB4
-50
[DXF | D
WG | 3D
CAD]


6

4

4

74

SRT
B450

Page 106

11.3

2D DRAWING RATCHET AND PAWL


Page 107


Page 108

12.0

FASTENERS AND BOLT

12.1

INTRODUCTION

1. Types of thread fasteners and gasket

Thread fasteners are reversible strengthening materials. This implies that they can be used and reused
again in a separate material that requires threading. There are several types of fastener threads. This
depends on the type and texture of the threads in the fasteners. You use fastener threads because they
are cost efficient and they are readily available in most tool shops and supermarkets. Most fasteners
require re-tightening as they, at times, fall off due to vibrations.

A thread is a ridge of uniform section in the form of a helix on the internal or external surface of a
cylinder (IFI description) or it could be described as a sloping plane curled around a cylinder.

External threads are on bolts or screws.
Internal threads are on nuts.
There are many forms of threads but two types are in common use on fasteners.
Machine Screw Threads - used on bolts, setscrews, machine screws and designed to mate with
preformed threads in nuts or tapped holes.
Exceptions may be thread forming screws like Taptite or self-drilling screws like Teksor thread cutters
like Type 23's, which form or cut their own machine screw thread.
Spaced Threads - used on woodscrews, self-tapping screws, coach screws and Type 25 thread cutters.
Designed to form its own thread, usually in a pre-drilled hole.
Exceptions may be self-piercing screws such as needle points or self-drilling screws like Type 17's
which create their own hole; some Teksmay also have spaced threads.


Page 109

There is a wide array of threaded fastener types, including:



Unified National Coarse Threads



UNC threads are the most common general fastener thread. Their fit is deeper and more
generic than that of a fine thread, allowing for easy removal. Generally, they have a
higher tolerance for manufacturing and plating, and do not need cross threading to
assemble.



Unified National Fine Threads



UNF threads have better torque-locking and load-carrying ability than UNC threads
because of their larger minor diameter. Because of their more specific fit, they have
tighter tolerances, finer tension adjustment, and can carry heavier loads. They are most
commonly found in the aerospace industry.



United National Extra Fine Threads



UNEF threads are finer than UNF threads; they are used in applications with tapped
holes in hard material, thin threaded walls, and tapped holes in thin material. As with
UNF threads, UNEF threads are common in the aerospace industry.



UNJC and UNJF Threads



There are two types of ―J‖ threads: external and internal. External UNJC and UNJF
threads have a larger root radius than the corresponding part (UNC, UNR, UNK, or UNF
threads). The larger root radius results in a larger tensile area than the corresponding
thread, and smaller stress concentration—bolts that carry heavy loads usually use ―J‖
threads.



UNR and UNK Threads



A UNR external thread is the same as a UNC thread, only the root radius is rounded.
There is no internal UNR thread. UNK threads resemble UNR threads, but the root radius
and minor diameter require inspection.



Constant-Pitch Threads



These threads come in a variety of diameters to fit a given application—bolts with
diameters of 1 in. and above commonly use pitches of 8, 12, or 16 threads per inch.


Page 110

2. Stress on threads, static loading on tensioned joint, gasket joint

The critical areas of stress of mating screw threads are


The effective cross section area, or tensile area, of the external thread.



The shear area of the external thread which depends upon minor diameter of the tapped
hole



The shear area of the internal thread which depends on the major diameter of the external
thread

The allowable stresses are reflected towards the surface of the material itself. We can see that it
is on full stresses as the shape started to change. From start, the material will started to
experience sheer stress, on that part, yield strength of the material can be return to its original
shape as it stores energy on the material itself, after a while, material will start to come into
maximum tensile limit where necking usually occur at that time. As time goes, ultimate tensile
stress will occur and the material will snap into 2 and that significant to failure.


Page 111

Stress, static and tension formulae.

D = Basic Diameter.
p = Screw Thread Pitch
Le = Length of Thread Engagement
A t = the screw thread tensile stress area
d p = Pitch circle diameter of thread
A ss =the thread shear area
The following formula for the Tensile Stress Area of the (male) screw

This is based on ISO 898 Part 1. See calculation below...
d p = Pitch circle diameter of thread
dp = (D - 0.64952.p )

The thread shear area = Ass
When the female and male threads are the same material.
Ass = 0.5. π. dp. Le = 0.5 π (D - 0.64952.p). Le
To ensure that the screw fails before the thread strips it is necessary the shear
area is at least 2 times the tensile area. i.e.


Page 112

Le (min) = 2 . A t / [0.5 .π .(D - 0.64952.p )]
This assumes that the male and female thread materials have the same
strength. If the Female Material strength is lower i.e J as calculated below is
greater than 1 then the length of engagement must be increased to prevent the
female thread stripping

If the value of J is greater than 1 then the length of engagement must be
increased to at least

Thus from above method of calculation, we can see that is proven shows on how the calculation
is been made from one parts to another. This can also be applied to other parts regarding the
shapes and sizes. For certain cases, we need to consider the material use for the parts as it
involved the modulus young’s of that material. Everything can be found in any references book
and it is helpful in solving any problems.


Page 113

3. Shear loading on bolted and riveted joints


The bolts are loaded in shear. Depending on the joint design the bolt can be in single or
double shear...



The bolt interface with the hole is compressively loaded. ( Crushing )

Single Shear..
Shear Stress = 4 . F / π. d 2
Compressive Stress = F / (d . t)
Plate Shear Stress = F / (2.c.t)
Double Shear..
Shear Stress = 2 . F / π. d 2
Compressive Stress = F / (d . t)
Plate Shear Stress = F / (2.c.t)
The stresses are adjusted based on the number of bolts / screws used for the
joint..


Page 114

Strength of bolts withstanding torsion generated shear loading

Consider a bracket taking an offset load F (N) at a radius R (m). The bracket
is secure using a number of bolts each with a Area A(m2 ). The bolts are
located around a centroid position each with a radius from the centroid of
rn(m) and a horizontal/vertical position relative to the centroid of hn /vn (m) . (
bolt is designated by the subscript "n". )

Location of Centroid...
The location of the centroid of the bolts can often be determined by inspection as in figure
above. If the bolts are not arranged around a convenient center then the centroid is determined
by..
X position = sum of the moments of area of all the holes about a fixed horizontal position
divided by the total hole area
y position = sum of the moments of area of all the holes about a fixed vertical position divided
by the total hole area


Page 115

The offset load is equivalent to a vertical force (F) + moment (F. R) at the
centroid of the bolts...

Each bolt is withstands a vertical shear force

Fnv = F / No of Bolts.

Each bolt also withstands a shear load

Fnm = F.R. rn / (r12 + r22...rn2)

The total horizontal force on each bolt

Fth= Fnm . vn / Sqrt(hn2 + vn2 )

The total vertical force on each bolt

Ft= Sqrt (Fth2 + Ftv2)

The total shear load on each bolt
The resulting bolt shear stress

Ftv= Fnv + Fnm . hn / Sqrt(hn2 + vn2 )

τ t = Ft /A

The shear stress in each bolt is calculated to ensure the design is safe..

Strength of bolt joints withstanding bending forces


Page 116

Each Bolt withstands a shear Force
The resulting shear bolt stress

Fs = Fv / (Number of bolts)

τ n = Fs /A

Note: Each bolt is assumed to withstand the same shear force.

If there are x bolts( numbered n = 1 to x). Then the tensile force withstood be
each bolt is designated Fnt i.e F1t,F2t, F3t....Fxt
A selected bolt (n) withstands a tensile force of
2

Vn / (V1 +

V22....Vx2

Fnt = ( Fv. Rv + Fh. Rh) .

)

The resulting tensile bolt stress

sn = Fnt /A

Maximum principals stresses in the bolt resulting from combined loading

The notes on this page Assuming all stresses developed only as a result of
bracket loading i.e zero preload and zero residual bolt torque...

Maximum principal tensile stress in the bolt

Maximum principal compressive stress in the bolt

Page 117

Maximum shear stress in the bolt

Failure criteria: Refer to page Failure Modes

The notes on this page In order to estimate the design factors of safety it is
necessary to consider the failure modes.

The preferred failure criteria for

ductile metals are the "Shear Strain Energy Theory" (Von Mises-Hencky
theory). For a stress regime associated with a bolt i.e pure tensile
stress sx combined with shear stress τ xy. The Factor of safety relative to the
material tensile strength Sy..Is calculated as follows

Factor of Safety = Sy / ( sx2 + 3 .τ xy2 ) 1/2


Page 118

4. Asymmetrical loading on bolted and riveted joints

Bolted and riveted connections are quite commonly encountered in various design problems.
Many trials have been carried out recently to determine the stress distribution in bolted and
riveted joints which is of extreme importance in the design of such joints. Effect of many
parameters has been investigated; however effect of thickness ratio of the joint plates did not
take too much attention. In this analytic work the effect of various plate thicknesses on the load
and stress distribution in cylindrical bolted joints has been investigated using the finite element
technique. Computed results show that the loads as well as the stress or pressure distribution
under bolt-head or nut are neither constant nor uniform. Whatever the thickness ratio of the two
plates (I1 /L2), constant load and stress occurred across the joint under the bolt-head end.
Maximum stress on surface and up to the mid plane level showed an increase with the thickness
ratio. However the maximum stress on interface decreases with the thickness ratio to reach its
lowest at L1 /L2 = 1. End of loading on surface or at mid plane increases with the increase in
thickness ratio, and tends to approach a constant value at L1 /L2 equals 10 or more—on interface
the opening position—represented by diameter of joint to bolt diameter D/d, reaches a maximum
value of 3.5 then decreases rapidly to reach a constant value of about 2.5 for L1 /L2 = 10 or
more. Also the thickness of the plate (L1) has a pronounced effect on the opening position.
The stress and shear in a rivet is analyzed like a bolted joint. However, it is not wise to combine
rivets with bolts and screws in the same joint. Rivets fill the hole where they are installed to
establish a very tight fit (often called interference fit). It is difficult or impossible to obtain such a
tight fit with other fasteners. The result is that rivets in the same joint with loose fasteners carry
more of the load—they are effectively stiffer. The rivet can then fail before it can redistribute
load to the other loose fit fasteners like bolts and screws. This often causes catastrophic failure of
the joint when the fasteners unzip. In general, a joint composed of similar fasteners is the most
efficient because all fasteners reach capacity simultaneously.


Page 119

5. Strength of welded joints

An accurate theoretical solution to force and strength conditions is an extremely complicated
problem for welded connections, even for welds with simple shapes. That is why common
technical calculations are based on a range of conventions and simplified premises. In view of
the strength checks, welded parts are usually considered a single compact part with a dangerous
spot (section) in the welded area. On the grounds that there is an even distribution of stress in the
active weld section, only theoretical rated stress in the specified section is specified for the
respective load, regardless of the technological workmanship of the weld or potential internal
tension. For connections with multiple welds, an even load on individual welds is assumed.
The strength checks of the connection are performed by simple comparison of the calculated
rated stress with the permissible stress in the weld. Permissible weld stress "SwA" is usually
specified from the value of the yield strength of the basic material "Re" based on the required
safety.


Page 120

6. Symmetrical loading on welded joints, brazing and soldering

Metalworking is the process of working with metals to create individual parts, assemblies, or
large-scale structures. The term covers a wide range of work from large ships and bridges to
precise engine parts and delicate jewelry. It therefore includes a correspondingly wide range of
skills, processes, and tools.
Metalworking is a science, art, hobby, industry and trade. Its historical roots span cultures,
civilizations, and millennia. Metalworking has evolved from the discovery
of smelting various ores, producing malleable and ductile metal useful for tools and adornments.
Modern metalworking processes, though diverse and specialized, can be categorized as forming,
cutting, or joining processes. Today's machine shop includes a number of machine tools capable
of creating a precise, useful work piece.
Here are some of the method of calculating load:

Load Rated stress [MPa, psi] Tensile/Press.

Shear

Bend


Page 121

Bend

Twist

Tensile

Tensile/Press.

Shear

Bend

Page 122

Tensile/Press.

Shear

Bend

Twist


Page 123

13.0

ASSEMBLY PART

Figure 13.1: assembly 2D and 3D drawing

Page 124

14.0

SIMULATION

14.1

INTRODUCTION
Simulation is the imitation of the operation of a real-world process or system over time.

The act of simulating something first requires that a model be developed; this model represents
the key characteristics or behaviors/functions of the selected physical or abstract system or
process. The model represents the system itself, whereas the simulation represents the operation
of the system over time.
Simulation is used in many contexts, such as simulation of technology for performance
optimization, safety engineering, testing, training, education, and video games. Often, computer
experiments are used to study simulation models. Simulation is also used with scientific
modeling of natural systems or human systems to gain insight into their functioning. Simulation
can be used to show the eventual real effects of alternative conditions and courses of action.
Simulation is also used when the real system cannot be engaged, because it may not be
accessible, or it may be dangerous or unacceptable to engage, or it is being designed but not yet
built, or it may simply not exist.
In this simulation for the one of design report requirement, we are focusing on ratchet and
pawl. The load that wills the ratchet accounted is 150N, by fixed a axis nodal at center of pawl.
All of this step are use simulation that known as ABAQUS.


Page 125

14.2

PICTURES

Figure 14.1 : ABAQUS simulation

Page 126

Figure 14.2 : full view of ratchet simulation

Page 127


Figure 14.3 : closed up simulation

Page 128


15.0

REFERENCES

15.1

WINCH


General information for winch: http://en.wikipedia.org/wiki/Winch



Design selection are from this website
http://www.etrailer.com/Winches/Fulton/FT20250301.html



http://www.pacifichoists.com.au/our-products/categories/winches/handwinches/pacific-brake-winches/



http://www.pacificmarine.net/marine-deck/winches-and-hoists/how-to-size-awinch.htm

15.2

COMPONENT


http://www.facebook.com/l.php?u=http%3A%2F%2Fwww.wirerope.com%2FPD
F%2Fsurelift_ctg_s.pdf&h=FAQHeW3ET


15.3

http://www.h-lift.com/eyehook.htm

DRUM



15.4

http://en.wikipedia.org/wiki/Winch
http://www.ingersollrandproducts.com/lifting/winches/drum.htm

GEAR


http://www.roymech.co.uk/Useful_Tables/Drive/Gears.html#Module



http://www.scribd.com/doc/15742009/DESIGNING-GEAR-IN-CATIA-V5R14HOWTO

15.5

HANDLE


http://www.westmarine.com/webapp/wcs/stores/servlet/WestAdvisorView?langId
=-1&storeId=11151&catalogId=10001&page=Winch Handles#.Uq9jSvQW1mw



http://en.wikipedia.org/wiki/Winch



http://www.freepatentsonline.com/4883255.html


Page 129

15.6

SHAFT


Dr. Mohd Idris Shah Bin Ismail, lecturer EMM 3506 Engineering Design I,
Mechanical and Manufacturing Department, Faculty Engineering, UPM



EMM 3506 Shaft Design Lecture Note by Dr. MohdIdris Shah Bin Ismail



EMM 3506 Different Types of Loading Lecture Note by Dr. MohdIdris Shah Bin
Ismail



S45C mild steel overview, Retrieved on 6th December 2013, 8.40pm, from
http://www.meadinfo.org/2010/03/s45c-jis-mechanical-properties.html



Shaft

key

design,

Retrieved

on

7

December

2013,

12.13am,

from

http://engineering.union.edu/~tchakoa/mer419/MER419_keys-and_couplings.pdf


Shaft Key cross sectional dimensions, Retrieved on 8th December 3.11pm, from
http://www.roymech.co.uk/Useful_Tables/Keyways/keyways.htm

15.7

LUBRICANT



http://en.wikipedia.org/wiki/Plain_bearing#Lubrication


15.8

http://en.wikipedia.org/wiki/Bearing_(mechanical)#Maintenance_and_lubrication

http://www.mecheng.iisc.ernet.in/~bobji/funtri/assign/Lubricants.htm

BEARING


http://www.rbcbearings.com/bushings/



http://www.nmbtc.com/bearings/engineering/design-shaft-housing.html


Page 130

15.9

FRAME


http://en.wikipedia.org/wiki/Stainless_steel



http://asm.matweb.com/search/SpecificMaterial.asp?bassnum=MQ304A

15.10 RATCHET AND PAWL




Pawl and Ratchet, Kohara Gear Industry Co. Ltd., retrieved
http://www.qtcgears.com/khk/newgears/khk316.html
Ratchets·Pawls, Kohara Gear Industry Co. Ltd
Model:
N01
Ratchet
and
Pawl
Mechanism,
retrieved
http://kmoddl.library.cornell.edu/model.php?m=247&movie=hide

from

from

15.11 FASTENERS AND BOLTS


http://www.ehow.com/info_8630066_types-fastener-thread.html



http://www.roymech.co.uk/Useful_Tables/Screws/Thread_Calcs.html



http://www.mitcalc.com/doc/welding/help/en/welding.htm



http://connection.ebscohost.com/c/articles/45037993/strength-stressed-stateasymmetric-mechanically-inhomogeneous-welded-joints-incomplete-fusioncenter-weld-under-biaxial-loading



http://avstop.com/ac/flighttrainghandbook/asymmetricloading.html



http://www.iso.org/iso/catalogue_detail.htm?csnumber=4736



https://www.excelcalcs.com/site-news/excelcalcs-news/built%11in-beam-with-2symmetric-point-loads/



http://en.wikipedia.org/wiki/Metalworking



http://www.slideshare.net/donalsyahrial/asm-metals-handbook-volume-6welding-brazing-and-soldering



http://www.aws.org/w/a/associations/bsmc/brazing_qa.html


Page 131

16.0
APPENDIX


Page 132

introduction, drawing, calculation for winch design

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