Fanstrap in Pulley Calculator - Centers RPM Belt Length and Speed , Gear , and Automotive industry AMNIMARJESLOW AL DO FOUR DO AL NON ZERO LJBUSAF thankyume orbit

                     Fan strap in Pulley Calculator - Centers RPM Belt Length and Speed



                                                              

Measuring Belts

How To Measure Standard Belts

 

Belts are NOT measured by diameter nor folding the belt in half. The most accurate way to measure the circumference of a belt is with a v belt measuring tool or using a cloth tape measure (Not a steel tape measure or ruler). An additional option is to wrap a small diameter string/rope around the belt, then measure the length of the string. Keep in mind belts do stretch and unfortunately there is no specific given value for how much a belt stretches.

Widths for standard A,B,C,D,E,AX,BX & A, B Kevlar Section V Belts:

Width's:

A Section=1/2" Wide

B Section=5/8" Wide

C Section=7/8" Wide

D Section=1-1/4" Wide

E Section=1-1/8" Wide

AX Section= 1/2" Wide

BX Section= 5/8" Wide

Widths for standard 3L,3V,5V,8V,3VX,5VX & 3L Kevlar V-Belts:

Width's:

3L= 3/8" Wide

3V= 3/8" Wide

5V= 5/8" Wide

8V= 1" Wide

3VX= 3/8" Wide

5VX= 5/8" Wide

How belt numbers are coded

All A,B,C,D,E,AX,BX belts have the inside circumference measurement after the letter of the part number. Example: A88 has a 88" inner circumference, then add 2 inches for the A Section (Listed below) to obtain the outer circumference. If you measured the outer circumference, then you would subtract the number of inches based on the section you have (Listed below) to obtain the belt number. This is also the same for the banded belts. Measure each individual band for banded belts to obtain the correct width.

Examples of how belt numbers are coded per section:

A Section= Add 2" to belt number (Example A88 = 90" Outside Circumference)

B Section= Add 3" to belt number (Example B88 = 91" Outside Circumference)

C Section= Add 4" to belt number (Example C88 = 92" Outside Circumference)

D Section= Add 5" to belt number (Example D105 = 110" Outside Circumference)

E Section= Add 6" to belt number (Example E144 = 150" Outside Circumference)

AX Section= Add 2" to belt number (Example AX60 = 62" Outside Circumference)

BX Section= Add 3" to belt number (Example BX50 = 53" Outside Circumference)

All sections 3L,3V,5V,8V,3VX,5VX belts have the outside circumference after the number and letter(s). Inside circumference are not measured for these sections. This is also for the banded belts. Measure each individual band for banded belts to obtain the correct width.

Examples of how belt numbers are coded per section:

3L Section= (Example 3L450 = 45" Outside Circumference)

3V Section= (Example 3V800 = 80" Outside Circumference)

5V Section= (Example 5V1180 = 118" Outside Circumference)

8V Section= (Example 8V2120 = 212" Outside Circumference)

3VX Section= (Example 3VX520 = 52" Outside Circumference)

5VX Section= (Example 5VX810 = 81" Outside Circumference)

Measuring K,L,M Micro Rib Belts:

Width's

K Section= 0.140" X's the number of ribs

L Section= 0.185" X's the number of ribs

M Section= 0.370" X's the number of ribs

All sections K,L,M belts will have the outside circumference in front of the letter. The rib count will be the number after the letter.

Outside Circumference & Rib Count per section:

K Section= (Example 845K8 = 84.5" Outside Circumference and 8 Ribs)

L Section= (Example 540L10 = 54" Outside Circumference and 10 Ribs)

M Section= (Example 940M6 = 94" Outside Circumference and 6 Ribs)  

 

 

V Belt Specifications

Classic V Belts - FHP (Fractional Horse Power)
Classic v belts are the most common belts used in the many industries across the world. All classic v belts are heat, oil and anti-static resistant.

3L Series a: 3/8" b: 7/32"

A Series (4L Series) a:1/2" b: 5/16"

B Series (5L Series) a: 5/8" b: 13/32"

C Series a: 7/8" b: 17/32"

D Series a: 1-1/4" b: 3/4"

E Series a: 1-1/2" b: 29/32"

Kevlar V Belts
Kevlar v belts are commonly used in the mower industry or any drives that require reverse bending, twisting or clutching. Kevlar v belts are heat, oil and anti-static resistant.

3L Series a: 3/8" b: 7/32"

A Series (4L Series) a:1/2" b: 5/16"

B Series (5L Series) a: 5/8" b: 13/32"

C Series a: 7/8" b: 17/32"

 

Classic Cogged V Belts

Cogged v belts transmit higher horse power ratings with improved heat dissipation. Cogged belts have greater flexibility than standard belts, which works better with smaller diameter pulleys. Cogged v belts are heat, oil and anti-static resistant.

AX Series a: 1/2" b:5/16"

BX Series a: 5/8" b: 13/32"

CX Series a: 7/8" b: 17/32"

Wedge V Belts

Wedge v belts generate higher horse power ratings with limited space where larger belts wouldn't work. Wedge v belts are heat, oil and anti-static resistant.

3V Series a: 3/8" b: 21/64"

5V Series a: 5/8" b: 35/64"

8V Series a: 1" b: 7/8"

Cogged Wedge V Belts

Wedge v belts generate higher horse power ratings with limited space where larger belts wouldn't work. Wedge v belts are heat, oil and anti-static resistant.

3VX Series a: 3/8" b: 21/64"

5VX Series a: 5/8" b: 35/64"

 

Metric V Belts

Metric v belts are similar to classic v belts except for these are manufactured to metric (British Standard) dimensions. Metric v belts are heat, oil and anti-static resistant.

SPZ Series a: 10mm b: 8mm

SPA Series a: 13mm b: 10mm

SPB Series a: 17mm b: 14mm

SPC Series a: 22mm b: 18mm

Double Angle V Belts

Double angle v belts are mainly used in applications where the belt is needed to reverse bend. Double angle v belts are heat, oil and anti-static resistant.

AA Series a: 1/2" b: 13/32"

BB Series a: 5/8" b: 17/32"

CC Series a: 7/8" b: 23/32"

 

Classic Banded V Belts

Classic banded v belts enhances power transmission capacity up to 25% and offers a permanent matched set. Classic banded v belts are heat, oil and anti-static resistant.

A Series a: 1/2" Per Band b:5/16" w/o band

B Series a: 5/8" Per Band b: 13/32" w/o band

C Series a: 7/8" Per Band b: 17/32" w/o band

D Series a: -1/4" Per Band b: 3/4" w/o band

 

Classic Cogged Banded V Belts

Classic banded v belts enhances power transmission capacity up to 25% and offers a permanent matched set. Great flexibility with improved heat dissipation. Classic cogged banded v belts are heat, oil and anti-static resistant.

BX Series a: 5/8" Per Band b: 13/32" w/o band

CX Series a: 7/8" Per Band b: 17/32" w/o band

 

Classic Wedge Banded V Belts

Wedge banded v belts enhances power transmission capacity up to 25% and offers a permanent matched set. Wedge banded v belts are heat, oil and anti-static resistant.

3V Series a: 1/2" Per Band b: 21/64" w/o band

5V Series a: 5/8" Per Band b: 35/64" w/o band

8V Series a: 7/8" Per Band b: 7/8" w/o band

 

Classic Cogged Wedge Banded V Belts

Wedge banded v belts enhances power transmission capacity up to 25% and offers a permanent matched set. Great flexibility with improved heat dissipation. Wedge banded v belts are heat, oil and anti-static resistant.

3VX Series a: 1/2" Per Band b: 21/64" w/o band

5VX Series a: 5/8" Per Band b: 35/64" w/o band

 

Micro Rib Belts
Extremely flexible and capable to work on small pulley diameters and serpentine drives. Micro rib belts are heat, oil and anti-static resistant.

K Series a: 0.140" Per Rib b: 0.177"

L Series a: 0.185" Per Rib b: 0.380"

M Series a: 0.370" Per Rib b: 0.660"

 

  V Belt Cross References For:

• Air Compressors
• Backhoes
• Banded V Belts
• Cars & Trucks
• Combines
• Chippers
• Fans
• Go-Karts
• Golf Carts
• Kevlar V Belts
• Riding Lawn Mowers
• Rock Crushers
• Round Balers
• Serpentine Drives
• Skid Steers
• Snow Blowers/Throwers
• Square Balers
• Tillers
• Farm Tractors
• Trenchers
• V Belt Distributors
• V Belt Manufacturers
• V Belt Suppliers
• Walk Behind Lawn Mowers
• Washer & Dryers
• Wedge V Belts
• Zero Turn Lawn Mowers
• And the list goes on...  

 

Kevlar K-Plus V-Belts

Kevlar V Belts

Kevlar/Aramid corded v belts have a higher tensile strength rating over standard v belts which have polyester cords.  This will minimize stretch when these belts are endure higher shock loads for longer periods of time.  Kevlar belts have a special dry cover/wrap for clutching applications and these can withstand reverse bending and twisting.


Benefits of the Kevlar (Aramid) V Belts:

• Excellent performance for drives having clutching applications and backside idler drives
• Provides reduced slippage while clutching
• Able to withstand high level of reverse flexing
• High resistance to larger shock loads (High tensile strength)
• Kevlar belts are great for the lawn mower and snow blower applications
• Superior resistance to oil, heat, cracking, ozone sunlight and weather

Sub-Categories

• Kevlar A Section K-Plus V-Belts

  1/2" Width
• Kevlar B Section K-Plus V-Belts

  5/8" Width

• Kevlar C Section K-Plus V-Belts

  7/8" Width
• Kevlar 3L Section K-Plus V-Belts

  3/8" Width

Applications:

•  Garden tractors
•  Riding lawn mowers
•  Commercial mowers
•  Walk behind lawn mowers
•  Blowers
•  Compost equipment
•  Snow throwers
•  Tillers
•  Chippers / Shredders
•  Sprayers
•  Lawn edgers
•  And much more...

Classic A,B,C,D,E V-Belts

Features

• Designed specially for reverse-bend drive by positioning tensile members closer to the bottom compare to that of conventional multiple V-belts A, B, & C.
• Particularly suitable to Agricultural & Industrial Machinery, where V belts are often driven with tension idler pulley on their back.
• Heat, Oil and Anti-Static resistant

Sub-Categories

• Classic A Section V-Belt

  1/2" Width
• Classic B Section V-Belt

  21/32" Width

• Classic C Section V-Belt

  7/8" Width
• Classic D Section V-Belt

  1-1/4" Width

• Classic E Section V-Belt                   
• 1-1/2" Width  

 

Classic Banded A,B,C,D V-Belts

Classic A, B, C, D Series Banded V Belts.

•  Enhanced power transmission capacity up to 25%
•  No lateral whip, twist, spin or jump-off
•  High adhesion strength between tie-band and the belts
•  Eliminates changes of mismatch in length (Permanent Matched Set)
•  Antistatic, oil and heat resistant
•  Manufactured through single-stage curing process

Sub-Categories

• Classic Banded A Section V-Belts

  1/2" Wide Per Band
• Classic Banded B Section V-Belts

  5/8" Wide Per Band
• Classic Banded C Section V-Belts

  7/8" Wide Per Band

• Classic Banded D Section V-Belts

  1-1/4" Wide Per Band

Banded belts are recommended under following conditions
•  Severe vibrations
•  Vertical shaft drives
•  V-flat drives
•  Agricultural drives
•  Conveyer systems
•  Stone crushers, reciprocating compressors, generator sets
•  And much more...

 

Classic Cogged AX,BX,CX V-Belts

Classic Cogged AX, BX, CX Section V Belts
•  Special top layer with fabric
•  Tough tensile member for greater strength
•  Molded Cogs
•  Improved heat dissipation
•  Greater flexibility for smaller diameter pulleys
•  Transmits higher horsepower
•  Antistatic, oil and heat resistant

Sub-Categories

• Classic Cogged AX V-Belts

  1/2" Width
• Classic Cogged BX V-Belts

  21/32" Width

• Classic Cogged CX V-Belts

  7/8" Width  

 

 

 

Classic Cogged Banded BX,CX V-Belts

•  Enhanced power transmission capacity
•  No lateral whip, twist, spin or jump-off
•  High adhesion strength between tie-band and the belts
•  Eliminates changes of mismatch in length (Permanent Matched Set)
•  Molded Cogs
•  Improved heat dissipation
•  Greater flexibility for smaller diameter pulleys
•  Transmits higher horsepower
•  Anti-static, oil and heat resistant

Sub-Categories

• Classic Cogged Banded BX V-Belts

  5/8" Wide Per Band
• Classic Cogged Banded CX V-Belts

  7/8" Wide Per Band  

 

Classic Wedge 3V,5V,8V V-Belts

•  Temperature Range: -18ºC to +80ºC
•  Maximum recommended belt speed is 30 m/s*
•  Maximum flex rate is f=100 per second
•  Antistatic, Oil and Heat resistant

Sub-Categories

• Classic Wedge 3V V-Belts

  3/8" Wide
• Classic Wedge 5V V-Belts

  5/8" Wide

• Classic Wedge 8V V-Belts

  1" Wide 

 

Classic Cogged Wedge 3VX,5VX V-Belts

•  Special top layer with fabric
•  Tough tensile member for greater strength
•  Molded Cogs
•  Improved heat dissipation
•  Greater flexibility for smaller diameter pulleys
•  Transmits higher horsepower
•  Antistatic, oil and heat resistant

Sub-Categories

• Classic Cogged Wedge 3VX V-Belts
• Classic Cogged Wedge 5VX V-Belts

  5/8" Width  

Classic Cogged Wedge Banded 3VX,5VX V-Belts

•  Enhanced power transmission capacity
•  No lateral whip, twist, spin or jump-off
•  High adhesion strength between tie-band and the belts
•  Eliminates changes of mismatch in length (Permanent Matched Set)
•  Molded Cogs
•  Improved heat dissipation
•  Greater flexibility for smaller diameter pulleys
•  Transmits higher horsepower
•  Antistatic, oil and heat resistant

Sub-Categories

• Classic Cogged Wedge Banded 3VX V-Belts

  3/8" Wide
• Classic Cogged Wedge Banded 5VX V-Belts

  5/8" Wide  

 

Classic Double AA,BB,CC Belts

•  Transmits power from both sides
•  Great flexibility
•  Can be driven clockwise and anti-clockwise direction simultaneously
•  Anti-static, Oil and heat resistant

Sub-Categories

• Classic Double AA Belts

  1/2" Width
• Classic Double BB Belts

  5/8" Width

• Classic Double CC Belts

  7/8" Width  

 

Classic Fractional 3L V-Belts

•  Recommended in special applications such as V-flat drives
•  Particularly suitable to Agricultural & Industrial Machinery
•  FHP Fractional horse power v belts
•  Anti-static, Oil and heat resistant

Sub-Categories

• Classic Fractional 3L V-Belts

  3/8" Width  

 

Classic Wedge Banded 3V,5V,8V V-Belts

•  No lateral whip, twist, spin or jump-off
•  High adhesion strength between tie-band and the belts
•  Eliminates changes of mismatch in length (Permanent Matched Set)
•  Antistatic, oil and heat resistant

 

Sub-Categories

• Classic Wedge Banded 3V V-Belts

  3/8" Wide
• Classic Wedge Banded 5V V-Belts

  5/8" Wide
• Classic Wedge Banded 8V V-Belts

  1" Wide  

 

Metric SPZ, SPA, SPB, SPC V Belts

Metric v belts are similar to classic v belts, except they carry British standard specifications and a different angle.  Metric v belts are supplied in 10mm, 13mm, 17mm and 22mm widths. 

Sub-Categories

• Metric SPZ Section V Belt

  (10mm Wide)
• Metric SPA Section V Belt

  (13mm Wide)

• Metric SPB Section V Belt

  (17mm Wide)
• Metric SPC Section V Belt

  (22mm Wide)  

Multi Speed Belts

•  Excellent transverse rigidity and flexibility to prevent bucking at minimum diameter settings where belt stress is more
•  Firm gripping action in contact areas provide positive traction for precise & immediate response
•  High power transmission than regular range of belts
•  Longer service life
•  Facilitates smooth running without excessive vibrations
•  Specific belt design for maximum longitudinal flexibility  

 

Micro Rib Belts (Serpentine Belts)

• Trapezoid faced ribs on a fibre reinforced rubber matrix for higher power transmission offering good resistance to wear and tear, facilitating quiet running
•  Reduced vibrations, shock absorber, low stretch and an excellent behavior under heavy load conditions
•  Extremely flexible, capable to work on small pulley diameters and serpentine drives
•  Oil and heat resistant, longer service life, suitable for HEMM applications

K Section Micro Rib Belts

L Section Micro Rib Belts

M Section Micro Rib Belts

 

 

 

Specialty Belts Go-Kart,ATV,Golf Cart, Snowmobile

We offer quality Go-Kart, ATV, Snowmobile and Golf Cart v belts and great prices.

•  Manufactured to the OEM specifications
•  Most specialty belts are Kevlar reinforced. (Helps prevent stretch)  

 

 

Lawn & Garden Belts

Click Below on your Brand Of Equipment or Input your V-belt Part number in our search box located at the top of the page. We Offer Over 1,300 Replacement Lawn Mower, Snow Blower V Belts For Many Different Manufacturers.

•  Manufactured to the OEM specifications
•  Most of our specialty belts are Kevlar (Aramid) cord reinforced. (Helps prevent stretch under high shock loads)

 

Snow Blower/Thrower Belts

 

• Ariens / Gravely Snow Blower Belts
• Craftsman Snow Blower Belts

• Cub Cadet Snow Blower Belts
• Husqvarna Snow Blower Belts

• John Deere Snow Blower Belts
• MTD / Cub Cadet Snow Blower Belts

• Murray Snow Blower Belts
• Noma / AMF / Dynamark Snow Blower Belts

• Sears / Craftsman / AYP Snow Blower Belts
• Snapper Snow Blower Belts

• Toro / Wheel Horse Snow Blower Belts
• Troy Bilt Snow Blower Belts

We offer a large selection of replacement snow blower/thrower belts for many manufacturers. If you don't see your belt listed? Simply enter the manufacturers belt number in the search box above to see of we offer a cross reference under an alternative number. 

•  Manufactured to the OEM specifications
•  Most of our specialty belts are Kevlar (Aramid) cord reinforced. (Helps prevent stretch)  

 

 

Twist & Lock V Belts

Do have equipment that's costly whenever one goes down due to a broken belt?  Well the Twist N' Lock is a quick, easy and effective solution tp get your machine back up and running in minutes.  These belts are available in the 3L, A(4L), B(5L) and C sections and are sold in 25 foot lengths.

NOT recommended for Lawn mower, snow blower or clutching applications.  This is a suitable replacement for standard v belt drive systems

Sub-Categories

• 3L Twist N' Lock V Belts

  3/8" Wide
• A / 4L Twist N' Lock V Belts

  1/2" Wide

• B / 5L Twist N' Lock V Belts

  5/8" Wide
• C Twist N' Lock V Belts

  7/8" Wide

  

                            

 

 

 

                                                                                 X  .  I   

                                                                  Design & Technology 

                                                             Mechanisms  Pulley systems  

A mechanism is simply a device which takes an input motion and force, and outputs a different motion and force. The point of a mechanism is to make the job easier to do. The mechanisms most commonly used in mechanical systems are levers, linkages, cams, gears, and and pulleys.

Levers: 1

You need to know how to calculate the mechanical advantage obtained by using levers, the velocity ratio in levers and pulley systems, and gear ratio and output speed when using gears.
A lever is the simplest kind of mechanism. There are three different types of lever. Common examples of each type are the crowbar, the wheelbarrow and the pair of tweezers.
All levers are one of three types, usually called classes. The class of a lever depends on the relative position of the load, effort and fulcrum:

• The load is the object you are trying to move.
• The effort is the force applied to move the load.
• The fulcrum (or pivot) is the point where the load is pivoted.

Class 1 levers  

Class 1 levers

A class 1 lever has the load and the effort on opposite sides of the fulcrum, like a seesaw. Examples of a class-one lever are a pair of pliers and a crowbar.
For example, it would take a force of 500N to lift the load in the animation below. But using a lever - a rod with the fulcrum placed closer to the load than the point of effort - it only requires a force of 100N. 

Class 2 levers

A class 2 lever has the load and the effort on the same side of the fulcrum, with the load nearer the fulcrum. Examples of a class-two lever are a pair of nutcrackers or a wheelbarrow.
In the diagram, the wheel or fulcrum on the wheelbarrow is helping to share the weight of the load. This means that it takes less effort to move a load in a wheelbarrow than to carry it 

Mechanical advantage and velocity ratio

Class 1 and class 2 levers both provide mechanical advantage. This means that they allow you to move a large output load with a small effort. Load and effort are forces and are measured in Newtons (N). Mechanical advantage is calculated as follows:

Mechanical advantage = load ÷ effort

In the example above, where the load=500N and the effort=100N, the mechanical advantage would be:
500N ÷ 100N = 5

Velocity ratio

The mechanical advantage gained with class-one levers and class-two levers makes it seem like you are getting something for nothing: moving a large load with a small effort. The catch is that to make the effort smaller, you have to move a greater distance. In the first diagram the trade-off is that you need to push the lever down further to move the load up a smaller distance. This trade-off is calculated by the velocity ratio:

Velocity ratio = distance moved by effort ÷ distance moved by load 

Levers: 2

A class 3 lever does not have the mechanical advantage of class-one levers and class-two levers, so examples are less common. The effort and the load are both on the same side of the fulcrum, but the effort is closer to the fulcrum than the load, so more force is put in the effort than is applied to the load. These levers are good for grabbing something small, fiddly or dirty, or picking up something that could be squashed or broken if too much pressure is applied. The common example of class 3 levers is a pair of barbeque tongs or a pair of tweezers.

Learn the position of load, effort and fulcrum for each class of lever. Linking it to a practical example will help you to remember them  

Linkages

Linkages are mechanisms which allow force or motion to be directed where it is needed. Linkages can be used to change:

• The direction of motion
• The type of motion
• The size of a force

A linkage consists of a system of rods or other rigid materials connected by joints or pivots. The ability of each rod to move will be limited by moving and fixed pivots. The input at one end of the mechanical linkages will be different from the output, in place, speed, direction and other ways.

Reverse-motion linkage

A reverse-motion linkage changes the direction of motion. In the diagram below, note how the linkage looks a little like a "Z". See how the central rod moves around a central fixed pivot. By pulling (or pushing) the linkage in one direction, it creates an exact opposite motion in the other direction. If the fixed pivot was not central, it would create a larger or smaller motion in the opposite direction.

Parallel-motion linkage

A parallel-motion linkage creates an identical parallel motion. In the diagram below, note how the linkage looks a little like an "n". This time, it is the two side rods that move around two central fixed pivots, while the top of the "n" moves freely. By pulling (or pushing) the linkage in one direction, it creates an identical parallel motion at the other end of the linkage.

Bell-crank linkage

A bell-crank linkage changes the direction of movement through 90°. A bell-crank linkage tends to look a little like an "L" or, as shown in the diagram below, a mirror image of an "L". By pulling (or pushing) the linkage in one direction, it creates a similar motion at the other end of the linkage. For example, a bell-crank linkage could be used to turn a vertical movement into horizontal movement, as in a bicycle braking system.

Treadle linkage

A treadle linkage shows how linkages can be used to change one type of motion into another. In this case, the rotary motion of the cam moves a parallel-motion linkage. The parallel-motion linkage controls the identical side-to-side, or oscillating motion, of two windscreen wipers. 

Cams

A cam is a shaped piece of metal or plastic fixed to a rotating shaft. A cam mechanism has three parts: cam, slide and follower.
The cam shaft rotates continually, turning the cam. The follower is a rod that rests on the edge of the turning cam. The follower is free to move up and down, but is prevented from moving from side to side by a slide or guide, so the follower can only do three things:

1. Rise (move up)
2. Fall (move down) or
3. Dwell (remain stationary)

The follower's pattern of movement depends on the profile or outside edge of the cam that it follows. If the cam is perfectly round and the fixed shaft is in the centre of the cam, the follower will dwell. But if the cam is a different shape, and/or the shaft is not central, the follower will rise or fall. How often and how quickly the follower moves is determined by the shape of the cam and the position of the shaft.

• Cams come in many different shapes - for example pear-shaped, triangular or square.
• The cam may have a chunk or chunks removed, so that the follower falls into a gap and is then is pushed out again.
• Whatever the shape of the cam, positioning the shaft off-centre will alter the behaviour of the follower.

Pear-shaped cam

The pear shape of this cam means that for half the cycle, the follower will dwell. Then, as the pointed part of the cam approaches, the follower is pushed up (rises), then, as the point passes, falls and dwells - and the cycle starts again.

Eccentric cam

The eccentric cam is perfectly circular, but the rotating shaft is off-centre, which affects how it turns. This type of cam produces a smooth, symmetrical rise and fall motion in the follower, which never pauses to dwell.

Drop cam

With a drop cam the shaft is central in a perfectly round cam, which has a chunk removed. The follower will dwell for most of the cycle, until it suddenly falls into the removed section, then rises again as the cam regains its circular shape.

Note that if the size of the cut-out portion was larger or the incline smoother, the follower would behave differently.

Gears: 1

Gears consist of toothed wheels fixed to shafts. The teeth interlock with each other, and as the first shaft (the driver shaft) rotates, the motion is transmitted to the second or driven shaft. The motion output at the driven shaft will be different from the motion input at the driver shaft - in place, speed, direction and other ways.
A number of gears connected together are called a gear train. The input (eg a motor) is connected to the driver gear. The output, (eg the wheel of a buggy) is connected to the driven gear.

Spur gears

The photograph below shows a simple gear train made up of a couple of spur gears. These are the common gears (or cogs) that look like wheels with teeth around the rim. Next to it is a diagram showing how you would draw this gear train in an exam.
In the drawing, the centre of each gear is shown by a cross. Each gear is drawn as two circles, one slightly larger than the other to show where the teeth would be. Teeth do not have to be drawn, but the number of teeth is written next to the gear, in this case 60 teeth and 15 teeth. Arrows indicate the direction that the gears are moving. Note that with two connected gears, they will be rotating in opposite directions.

Gear ratio and output speed

Where there are two gears of different sizes, the smaller gear will rotate faster than the larger gear. The difference between these two speeds is called the velocity ratio, or the gear ratio, and can be calculated using the number of teeth. The formula is:

Gear ratio = number or teeth on driven gear ÷ number of teeth on the driver gear

So the gear ratio for the simple gear train above, if the smaller gear is the driver gear, is:
Gear ratio = 60 ÷ 15 = 4.
In other words, the driver gear revolves four times to make the driven gear revolve once.
If you know the gear ratio, and the speed input at the driver gear, you can calculate the speed output at the driven gear using the formula:

Output speed = input speed ÷ gear ratio

So if the gear ratio is 4 and the driver gear is revolving at 200 rpm then the output speed = 200 ÷ 4 = 50 rpm  . 

Gears: 2

Compound gear train

Where very large speed reductions are required, several pairs of gears can be used in a compound gear train. A small gear drives a large gear. The large gear has a smaller gear on the same shaft. This smaller gear drives a large gear. With each transfer, the speed is significantly reduced.

Worm gears

Another method of making large speed reductions is to use a worm gear. This is a shaft with a thread like a screw. This connects at 90° to a large gear (the thread shaft points along the outside edge of the larger gear). Each time the shaft spins one revolution, the gear turns forward by only one tooth. If the gear has 50 teeth, this creates a gear ratio of 50:1. The worm can drive the worm gear round, but the worm gear cannot drive the worm. This means that worm gears are good to use in hoists, the load will not fall back when the motor stops. Worm gears are a good option when you wish to alter direction or rotary motion through 90° and reduce the speed. The photograph to the left shows a worm gear powered by a motor.

Bevel gears

Bevel gears, like worm gears, change the axis of rotation through 90°. The teeth have been specially cut so the gears will mesh at right-angles to each other, where spur gears must be parallel.

Rack and pinion

A pinion is a round cog and the rack is a flat bar with teeth. The driver cog either moves along the rack, as in a rack and pinion (funicular) railway - or else the driver cog moves the rack, as in the steering system in cars .

  

Pulley systems

Pulleys are used to change the speed, direction of rotation, or turning force or torque.

A pulley system consists of two pulley wheels each on a shaft, connected by a belt. This transmits rotary motion and force from the input, or driver shaft, to the output, or driven shaft.

If the pulley wheels are different sizes, the smaller one will spin faster than the larger one. The difference in speed is called the velocity ratio. This is calculated using the formula:

Velocity ratio = diameter of the driven pulley ÷ diameter of the driver pulley

If you know the velocity ratio and the input speed of a pulley system, you can calculate the output speed using the formula:

Output speed = input speed ÷ velocity ratio

Worked example

• Work out the velocity ratio and the output speed of the pulley shown in the diagram above.
• Velocity ratio = 120mm ÷ 40mm = 3
• Output speed = 100rpm ÷ 3 = 33.3 rpm

Torque

The velocity ratio of a pulley system also determines the amount of turning force or torque transmitted from the driver pulley to the driven pulley. The formula is:

output torque = input torque × velocity ratio.

Pulley drive belts

Drive belts are usually made of synthetic fibres such as neoprene and polyurethane, with a V-shaped cross section. It is possible to reverse the direction of the driven pulley by twisting the belt as it crosses from input to output. Pulley belts have the advantage over chains that they do not need lubrication (though unlike a chain, a belt can slip).  

Types of motion

There are four basic types of motion in mechanical systems:

• Rotary motion is turning round in a circle, such as a wheel turning.
• Linear motion is moving in a straight line, such as on a paper trimmer.
• Reciprocating motion is moving backwards and forwards in a straight line, as in cutting with a saw.
• Oscillating motion is swinging from side to side, like a pendulum in a clock.

Many mechanisms take one type of input motion, and output it as a different type of motion. Below are some examples.

Chain and sprocket

A chain and sprocket changes rotary motion to linear motion - or vice versa. A wheel-and-axle, rack-and-pinion, rope-and-pulley, screw thread, or chain-and-sprocket could also be used for this.

Cam-and-follower

A cam-and-follower changes rotary motion to reciprocating motion.

Peg-and-slot

A peg-and-slot changes oscillating motion to rotary motion. A crank, link and slider could be used for this, and also to change rotary to oscillating motion.

Crank, link and slider

 A crank, link and slider will change rotary motion to oscillating / reciprocating motion.

Rack-and-pinion

A rack-and-pinion changes rotary motion to reciprocating motion. A crank, link and slider could also be used for this. A cam-and-follower will change reciprocating to rotary motion.

Mechanical systems and sub-systems

Small systems can be combined to make more complex systems. A cam which is turned by an electric motor can operate a micro switch which could be used to turn a light on or off. Two mechanical systems can be connected together to give complex movements. 


An Introduction to Variable Speed Pulleys and How They Work

A Variable Speed Pulley is a device for changing the speed of a rotating shaft member, so that it will increase or decrease the revolutions per minute of that particular shaft. The belt used in the pulley has a 'V' cross section, where the sides of the 'V' conform to the angular faces of the pulley flanges. As the flanges of the pulley open or close, the belt will be forced either up or down in the pulley, producing a larger or smaller working pitch diameter which in turn will increase or decrease the revolutions per minute of the shaft upon which the pulley is mounted. Usually two pulleys are mounted on separate shafts, one called the driver shaft which is always the power source, such as an electric motor, the other called the driven shaft which is the unit where the speed variation is required. The belt is simply the medium used to transfer the power from the power source to the driven unit. The power transmitting capacity of the belt is achieved by frictional grip, which is the wedge action of the belt between the pulley faces.

There are basically two types of systems in use:

The Adjustable Center Type Drive The adjustable pulley is spring loaded, the spring being the medium used to keep tension on the belt for maintaining the frictional grip. As the pulley, mounted on the motor shaft is forced away from the driven fixed pitch sheave by the screw action of the adjustable motor base, the belt is forced to ride on an ever decreasing pitch diameter, constantly slowing the speed of the driven pulley, until the required speed is reached.

The Fixed Center Type Drive Here, both pulleys are adjustable, the driven pulley being spring loaded, the driver pulley mechanically variable by means of an adjusting screw and nut. The action of turning the screw in the driver pulley will open the faces, allowing the belt to ride lower between the vees. The spring in the driven pulley will then force the belt to ride higher in that pulley, as the flanges are forced closer together. This will decrease the speed of the driven pulley. When the action is reversed by turning the screw in the driver pulley in the opposite direction, thereby allowing the pulley faces to close, the reverse will apply, and the speed of the driven pulley will increase. In this type of drive, the driver and driven shafts are in a fixed position. They can only rotate, not move in any other direction.  

  

BELT DRIVES

POSITIONS OF SLACK AND TIGHT SIDES OF BELT

While the belt is running, the belt tension is such that ‘sag’ or ‘droop’ is visible on one side of the driving pulley. This is shown in Fig. 2.4.1for flat belt drive.                               Fig. 2.4.1 Open belt drive with slack side on top of pulleys   2.4.1 Moving open belt with slack side on top of pulleys The positions of input and output pulleys are such that the tight side of the belt must be on the bottom and slack side on the top of the pulleys. Otherwise, the angle of contact between the belt and rim of the pulley reduces, decreasing the power transmission capacity of the belt.                                                                                      X  .  II   DRIVES IN MACHINES Belt drives such as flat belt, V belt, round belt, timing belt and tape (thin belt made from cloth and composite) are widely used in textile machines. They are simple and inexpensive compared to gears drives. Belt drive requires an endless belt and two pulleys (a driver and driven). Mostly they are used to transmit power between two parallel shafts by means of friction. The belt must be set with some initial tension to avoid it slipping over the pulleys for effective power transmission. Depending on the cross-sectional shape of belts, they are classified as flat, V and round belts. Belt drives offer maximum versatility as power transmission elements. The designer has considerable flexibility in choosing the location of pulleys for the driver and driven. They are used for power transmission over comparatively long distances. The design tolerances for these drives are not as critical compared with gear drives. In many cases, their use simplifies the design of machine and substantially reduces the cost. The advantage with belt drives is that they reduce vibration and shock transmission, since the belts are elastic and usually quite long. These properties play an important part in absorbing shock loads and isolating the effects of vibration. This aspect is very important for the life of machine. The belt drives are relatively quiet. The movement of belt depends on friction traction on the pulleys and hence, some slippage is inherent in their operation. The slippage of belt over the pulleys is also responsible to absorb shocks and vibrations. Some slip and creep are inherent in flat and v-belts, and so the angular velocity ratio between the driver and driven is neither constant nor equal to the ratio of the pulley diameters. Due to ageing or creep of belts, in some cases, an idler or tension pulley must be used to avoid the adjustments in center distance between the driver and driven pulleys. The belts with excessive creep must be replaced with new belts. Periodic inspection of belt slackness is required. Belts do not have an indefinite life. While in use, it is essential to have regular inspection schedule to guard against wear, ageing and loss of elasticity due to creep, so that they can be replaced at the first sight of deterioration.     FLAT BELT DRIVE Flat belts have narrow rectangular cross-section. In fact the earliest belt used in industrial drives was leather flat belt. Larger flat belt drives were in use as group-drive system in industry decades ago. A large motor drives several machines through pulleys and leather belts. Later, reinforced flat flats were introduced which have almost replaced the leather belts due to their superior characteristics. The important material properties to be considered for the construction of flat belts are high coefficient of friction between the belt and the rim of the pulleys, flexibility, durability and strength of the belt. Leather belts Leather belts were widely used earlier. Leather belts offer moderate coefficient of friction between the rim of the pulley and belt. But they become rigid, and exhibit creep over a period of time. They also have poor resistance against moisture. The leather belts are available in two varieties, viz., oak-tanned and mineral or chrome-tanned. Few layers of leathers are bonded together by adhesives to get the required thickness of belt. Commercial leather belts are specified according to the number of layers, such as single, two, three and four-ply belts. A three-ply leather belt is shown in Fig. 2.2.1. Fig. 2.2.1 Three-ply leather flat-belt Reinforced belt Another category of flat belts is the reinforced belts, which are widely used nowadays. The reinforced belts are made of urethane or rubber matrix, reinforced with fabric or nylon cords or steel wires. In the case of fabric reinforced rubber belts, canvas- or cotton-ducks fabrics are used. Rubber impregnated fabric belts are cheaper, have more resistant to moisture than leather belts. Either one or both the surfaces of belt (the later is used in reversing drives, discussed in section 2.8) require friction surface coating. Flat belts are quite, efficient at high speeds, and can transmit large amounts of power over long distances. They are mostly available in roll form and are cut into required length, the cut ends are joined using special kits. The coefficients of friction of leather, polyamide and urethane flat belt are 0.4, 0.5 to 0.8 and 0.7 respectively. The reinforced belts are available with a density of 0.97 to 1.29 g/cm3. The thickness of flat belts ranges from 0.75 to 5 mm. There is no upper limit on the length of flat belt to be used. Pulleys for flat belt The pulleys used for flat belt are crowned to keep the belts from running off the pulleys. Both the driver and driven pulleys must be crowned when the pulley axes are not in a horizontal position. The crowns should be rounded and not angled. If only one pulley is to be crowned, then it should be the larger one.     ANALYSIS OF FLAT BELT TENSIONS Consider a small element of a flat belt resting over a pulley shown in Fig. 2.3.1. Fig. 2.3.1 Forces acting on an element of a flat-belt The normal force (dN) acting on the belt arises due to reaction from the pulley. The coefficient of friction between pulley and belt is ‘μ’. Due to friction traction, the belt tensions on both the sides of the element are F1 (on loose side) and F2 (on tight side), such that F2 > F1 and the friction force dF is equal to difference between these two forces. The angle of wrap of belt over the pulley is (in radians). If a belt having a linear density of m (in kg/m) running over a rim of pulley at a velocity, v (in m/s), the element of belt shown in figure would be subjected centrifugal force equal to (in N) The above equation indicates that power transmission is proportional to belt speed. However, at very high belt speeds (usually above 1500 m/min), power decreases with increasing belt speed due to rapid rise of centrifugal force acting the belt. This centrifugal force reduces the pressure between the belt and the rim of the pulleys, moving the belt away from the pulley, reducing the wrap angle and hence, the belt tensions and power transmission.   MAXIMIZING THE POWER TRANSMISSION GEOMETRICAL RELATIOSHIPS IN BELT DRIVES Flat belts are used in open and crossed configurations. A crossed belt is shown in Fig. 2.6.1. The geometry of open flat belt and crossed flat belt drives are shown in Fig. 2.6.2. Fig. 2.6.1 Crossed belt on a high speed card Fig. 2.6.2 Geometry of flat-belt drives: Top-Open belt; Bottom- Crossed belt The contact angles (in degrees) of open belt over the smaller (driver) and larger (driven) pulleys (Fig. 2.6.2) are given below: SELECTION OF BELT AND PULLEY DIAMETER The machinery manufacturer has to select a belt from the belt manufacture’s catalogue based on the power to be transmitted, speeds and diameters of driving and driven pulleys and the available space to house the pulleys. In order to give factor of safety to the belt, the actual power to be transmitted by the belt is multiplied by load correction factor to arrive at maximum power transmission. The correction factor is one for a normal load and goes up to 1.5 when the nature of load is shock with increasing intensity. From the pulley manufacturer’s catalogue, a pulley whose size is nearest to the dimension as per the design calculation has to be selected. As a consequence the belt velocity and the size of the driven pulley would also vary slightly.       REVERSING DRIVES WITH FLAT BELTS Reversing drives (open or crossed) are used when the direction of rotation of driven must be opposite to that of driver. This can be achieved by open flat belt (Fig. 2.8.1) and crossed flat belt (Fig.2.8.2) configurations. In these configurations, both the sides of belt contact the pulleys. The V belts cannot be used for reversing drives. However, a double sided timing belt can be used in an open configuration (Refer Fig. 2.17.2). Fig. 2.8.1 Reversing drives with open flat belt Fig. 2.8.2 Reversing drives with crossed flat belt         FLAT BELT WITH OUT OF PLANE PULLEYS Flat belt drive could also be used when the pulleys are out-of-plane, which is not possible with V belt drive. The driving and driven pulleys must be positioned so that the belt leaves each pulley in the mid-plane of the face of other pulley without using a guide pulley. This drive is used to drive coiler plate on card as shown in Fig. 2.9.1.The schematic representation of the same is illustrated in Fig 2.9.2. For other arrangements guide pulleys are needed. Fig. 2.9.1 Flat belt on out-of-plane pulleys driving coiler on carding machine Fig. 2.9.2 Quarter-twist flat-belt drive (Pulleys are out-of-plane by 90°)      CLUTCHING ACTION WITH FLAT BELT Clutching action can be obtained with flat belts. The belt can be shifted from a loose to a fast (tight) pulley mounted on the driven-shaft using a fork-lever or any other shifting mechanisms as shown in Fig. 2.10.1. This type of arrangement is very common on conventional blow rooms, drawing and roving machines. While starting the motor, the belt is placed on the loose pulley, which disconnects the motor from machine, thus, preventing transfer of heavy loads on the motor. When the motor attains full speed, the drive is transferred to the machine by shifting the belt on to the fast pulley. Clutching device based on loose and fast pulleys used on a old carding machine is shown in Fig. 2.10.2. Fig. 2.10.1 Flat belt on fast and loose pulleys   Fig. 2.10.1 (b) Flat belt on fast and loose pulleys mounted on cylinder shaft of a card             APPLICATIONS OF FLAT BELTS Some of the applications of flat belts are given below: Drives to beaters on conventional blow rooms. Crossed flat-belt transmits drives from cylinder to flat on old cards.   Drives in high production cards such as the drive from motor to lickerin and cylinder; drive to cleaner roller at the delivery side; drive from motor to flat-stripper roller and crossed-flat-belt drive from cylinder to a pulley from where further drive proceeds through double stage speed reduction using worm and worm gears and a mechanical clutch to the driving-shaft of flat. Drive to drafting rollers and other rolling elements on a single delivery drawing machine. Drives to opening rollers, friction drums and take-off rollers on friction spinning machine. Drive to rotor on rotor-spinning machine. Main drive on draw-texturing machine. Drive to creel-rollers of a high speed drawing machine.         BELT DRIVES The V belts are the probably the most common means of transmitting power between fractional horse power motors to machines (Fig 2.12.1). Mostly, the driver and driven pulleys lie in the same vertical plane. There is an upper limit on the center distance or belt length. Long center distances are not recommended, because the excessive vibration of slack side flutters and shorten the belt life. In general the center distance should not be greater than 3 times the sum of diameters of input and output pulleys. Since the V belt is short, it is subjected to the action of load and fatigue a greater number of times. Further, its ability in absorbing the shocks is poor. Fig. 2.12.1 (a) V belt drive on a laboratory model card Fig. 2.12.1 (b) V belt drive on a loom Fig. 2.12.1 (c) V belt drive on a roving machine The V belts work better in the speed range 300 to 1500 m/min. V-belts are widely used in variable speed drives using adjustable sheaves. By moving the sheaves axially the pitch diameters of the driving and driven pulleys could be varied to get variable output speed. This type of drive is common on ring spinning and rotor spinning machines. Quarter-turn drives are used to transmit motion between horizontal and vertical shafts using deep groove pulleys and relatively long center distances.     CONSTRUCTION OF V BELT V belts are available without any joints. The cross-section of V belt is trapezoidal (shown in Fig. 2.13.1 ). V beltsare designed to mesh in the trapezium shaped groove of pulleys. The groove angle (β) of the sheave is made somewhat less than the belt-section angle (θ). This causes the belt to wedge itself into the groove, thus increases the friction. This increases greatly the frictional resistance to slipping, for a given maximum tension, compared to flat belt. The V belt does not rest on the bottom of the groove but wedges itself into the groove. Fig. 2.13.1 Geometry of a V belt The thickness ‘b’ ranges about 8 to 23 mm and the width ‘a’ is about 13 to 38 mm. The belt section angle is generally around 40° . The V belts are available in sections as A, B, C, D, E. The width and thickness of belt and minimum sheave diameter increase from section A to E, in other words, belts become heavier from section A to E. To select a V belt for different speeds and power transmission combination, one must refer to belt manufacturer’s catalogue. For high power transmission, heavy V belts are preferable. For high speeds, light belts are preferred. V belts are made of fabric and cords which are moulded in rubber. It consists of three elements, viz., a central load carrying layer of cords made from high tenacity fibre or steel, a surrounding layer of rubber to transmit pressure from cords to side walls and an elastic outer cover to provide friction as shown in Fig. 2.13.2. Fig. 2.13.2 Cross-section of a V belt       FORCE ANALYSIS IN V BELT Fig. 2.14.1 Forces on a V belt     APPLICATIONS OF V BELTS   Few applications of V belts are listed below: Main drives (drive from motor) in all spinning, yarn preparatory, texturing machines, looms, warping and winding machines and compressors. Drive from top coiler to base coiler plate of high speed carding machine.    ROUND BELTS The round belts are circular in cross-section, without any joints. They can be used for transmission over a long distance. The pulleys on which it runs should have V-shaped groove, similar to the ones used for V belts. This belt could also be used for variable speed drives using stepped pulleys. When all the pulleys (driving, driven and guide pulleys) are at considerable distances, and at different planes, these belts are the ideal choice. One such example is the drive to grooved winding drums on friction-spinning machine. The drive to winding drum on a friction-spinning machine is shown in Fig. 2.16.1 and the schematic representation of the same is shown in Fig. 2.16.2. The motion is transmitted from a pulley, ‘A” to pulley, ‘B’ through guide pulleys, ‘C’ & ‘D’. From the pulley, ‘B’ drive is transmitted to a grooved winding drum (not shown in the picture). Fig. 2.16.1 Round belt drive on friction spinning machine Fig. 2.16.2 Line sketch of round belt drive on friction spinning machine       TOOTHED BELT DRIVES They are also referred as ‘timing belt drives’. They are positive drives that operate on toothed pulleys. The belts have flat outer surface and evenly spaced teeth on the inner surface. A toothed belt is made of rubberized fabric reinforced with steel wires to take the load. The steel wire is located at the pitch line and the pitch length is the same regardless of the thickness of belt. Toothed belts do not have joints. The toothed pulley looks like a spur gear (shown in Fig. 2.17.1), but the tooth profile is not involute. Fig. 2.17.1 Timing-belt drive Characteristics of toothed belt drive Toothed belts offer very good accuracy in transmitting motion compared to flat belts and are comparable to gears. In addition they offer greater flexibility in the location of driver and driven. The tensions on timing belts are low, consequently the load on the supporting bearing are also low. They are commonly used on high-speed machines, when the distance between driver and driven is considerably long. In these situations they offer greater advantage over gear drives in terms of lower power consumption and noise. In the case of gear drive, train of gears is required with several carrier gears, which, leads to high power consumption and more noise. In addition, the gear drive becomes so complex that changing the gears to alter process parameters or removing and refitting the gears during maintenance operation would be cumbersome. Toothed belt do not stretch or slip much, consequently transmits power at constant angular velocity. Timing belts do not require high initial tension required for the flat and V belts. They operate over a wide range of speeds, with efficiencies in the range of 97 to 99%. They are quieter than chain drives. There is no periodic speed variation, as with chain drives, and so they are an attractive solution for precision-drive requirements. There is an upper limit on the maximum center distance between the wheels, which is lower compared to the flat belts. Applications of toothed belts Toothed belts are used on texturing machines and to drive doffer, stripper, calender and coiler rollers (double sided toothed belt) on high speed card, and coiler plates of combing machine. In new drawing and ring spinning machines the drafting rollers are driven by toothed belt drives. The phenomenon of fluctuating speeds with gear drives due to accumulation of fibres/dust on gear teeth, wearing teeth, and improper meshing of gears by deflection of shaft or misalignment of gears are eliminated with the use of timing belt drives. In looms, toothed belt drives are used to drive take-up roller, dobby/tappet shaft and warp let-off motion. Toothed belt drives used on an air-jet texturing machine and card are shown in Fig. 2.17.2. Fig. 2.17.2 Timing belt drive on air-jet texturing and carding machines The driving and driven wheels have discs on their sides to prevent belt coming out, in case the wheels are misaligned. The belt is kept under tension by means of a tension wheel (Fig 2.17.3), which does not have side discs. This helps in sliding the belt over the tension pulley after changing the driver or driven pulleys. Fig. 2.17.3 Tension wheel on a timing belt drive   The driving and driven wheels have discs on their sides to prevent belt coming out, in case the wheels are misaligned. The belt is kept under tension by means of a tension wheel, which does not have side discs. This helps in sliding the belt over the tension pulley after changing the driver or driven pulleys.     TAPES Tapes are very thin and highly flexible compared to flat belts. They are generally made from narrow woven fabrics. Tapes are very useful to drive a group of elements from a single source, following very tortuous paths. They can easily follow sharp curved paths, and bend and twist over the supporting or tension compensating pulleys. Tapes are used in ‘Four spindle group drive system’ to drive spindles on ring spinning machines (shown in Fig. 2.18.1 ). Fig. 2.18.1 Four-spindle group-drive system with tape Each tape moves over a tin roller or jockey pulley of driving shaft and tension pulleys to drive a set of four spindles. The spindle wharves are crowned. Thin flat belts are also used as ‘Tangential drive’ to drive spindles.       VARIABLE SPEED DRIVES Cone and stepped pulleys For variable speed drives in blow rooms and roving machines, flat belts with cone pulleys are used as shown in Fig 2.19.1. The belt is moved axially to vary the output speed. For stepped pulleys, V belt or round belt is used with grooved sheaves as shown in Fig. 2.19.2. The stepped pulleys with V belts are commonly used on many main drives of textile machines. Click on Image to run the animation Animation 2.19.1 Variable speed using cone pulleys Fig. 2.19.1 Variable speed drive with cone pulleys Click on Image to run the animation Animation 2.19.2 Variable speeds with stepped pulleys Fig. 2.19.2 Variable speed drive with stepped pulleys Conical discs A variable speed drive using adjustable grooves/conical discs and V belt are commonly used in ring spinning without varying the speed of the motor as shown in Fig. 2.19.3 and Fig. 2.19.4. Fig. 2.19.3 Speed variation using conical discs on ring spinning machine Click on Image to run the animation Animation 2.19.4 Speed variations with conical discs Fig. 2.19.4 Speed control on ring spinning using conical discs By shifting the driver and driven discs axially and simultaneously, the effective diameters of the discs over which the belt passes are varied, thus varying the output speed. To increase the output speed (spindle speed), the input discs are moved closer to each other and the output discs are moved apart and vice versa. A microprocessor controls the hydraulic or pneumatic piston and lever mechanism to moves the discs. Depending on the preciousness of the control mechanism, the speed of the output can be varied infinitesimally and continuously. This is called PIV (Positively Infinitesimally Variable) drive. However, the spindle speed in ring spinning is not continuously varied. In practice, the spindle speed is varied in several steps depending on the doff-position and the permissible end-breakage rate of yarn. This permits higher throughput of yarn as optimum spindle speed could be selected at any instant. To reduce slip even further, the V-belts are replaced by a set of steel links held together by means of a chain (slated chains). This is called PIV gear and is used in many industrial machines for speed control.       ADJUSTMENT OF BELT TENSIONS Belts become slack due to creep during its service life. Therefore, a provision should be made to adjust the belt tension from time to time. Different methods are available to adjust belt tensions. In the case of flat belts with joints or hinges, a short length of belt is cut periodically and the cut ends are joined back to remove the slackness in the belts. Movable and swinging motors For endless belts such as V belts and some flat belts, the technique of cutting the belt to adjust belt tension is not possible. In such cases, the center distance between input and output pulleys is slightly increased by means of an adjusting screw. On the drive from motor to main shaft of machine, provision is made to move the motor away from the output pulley through adjusting screws on the motor platform as shown in Fig. 2.20.1. Fig. 2.20.1 (a) Motor mounted on movable flat bed on a loom Fig. 2.20.1 (b) Motor mounted on movable flat bed Another way of adjusting the belt tension is to mount the motor on a swinging platform/hanging plate, called ‘Rockwood belt drive’ or ‘Pivoted motor’ used in bale opener is shown in Fig. 2.20.2. This type of drive is used in bale opener to drive the take-off and evener rollers. The motor is mounted on an over-hang-plate. Fig. 2.20.2 Swing motor drives on Bale opener In pivoted motor drive, the belt length is adjusted automatically, when the distance ‘z’ of the center of gravity of motor from the pivot changes due to creep on the belt (Fig. 2.20.3). Hence, the belt tension is also adjusted. Taking moments of forces at the pivot, Fig. 2.20.3 Forces acting on a pivoted motor drive In this drive, the belt length is adjusted automatically, when the distance ‘z’ of the center of gravity of motor from the pivot changes due to creep on the belt. Hence, the belt tension is also adjusted. Taking moments of forces at the pivot, Idler or tension pulley Adjustment of belt tension can also be carried out using an idler pulley. This idler pulley may be either spring-loaded ( Fig. 2.20.4) or held against the belt by its own weight (Fig. 2.20.5) or by external adjustable tensioning arrangement (Fig. 2.20.6). Fig. 2.20.4 Spring loaded idler pulley on friction spinning machine   Fig. 2.20.5 Weighted idler pulley In all the cases, the idler pulley should be located next to driver pulley on the loose side of belt. The contacting face of the idler pulley is flat faced without any crown. The idler pulley increases the arc of contact between belt and driver pulley, which is also advantageous in drives with short center distance. A flat-belt drives lickerin and cylinder through adjustable tension pulley on a high speed carding machine ( Fig. 2.20.6). Fig. 2.20.6 Flat belt drive to lickerin and cylinder with tension pulley A tension pulley is placed on the loose side of belt close to the motor pulley. The tension pulley is mounted on a slot fixed to the machine frame. Belt tension is adjusted by moving the tension pulley through the slot. When the machine is not operated for long period, the tension on the belt can be released to avoid unnecessary creep on the belt and to improve the service life of belt.   X  .  III CHAIN DRIVES   INTRODUCTION A chain drive consists of an endless chain wrapped around sprocket wheels (shown in Fig. 3.1.1). The chain has a number of links connected by pins. The sprockets have teeth of special profile. Chains are used for power transmission and as conveyors. The chain drives have some features of both belt (flexibility of location of driver and driven) and gear drives (ruggedness). Chain drives are recommended for velocity ratio below 10:1, chain velocity 1550 m/min and power transmission up to 100 kW.   Fig. 3.1.1 Chain drive            CONSTRUCTION OF ROLLER CHAIN Roller chain is made up of alternate link plates (inner and outer), pins, bushes and rollers as shown in Fig. 3.2.1. Fig. 3.2.1 Construction of a roller chain The pins, bushes and rollers are made of alloy steels. The pins are press fitted to the outer link plates. The bushes are press fitted to two inner link-plates. The bush and the pin form a swivel joint and the outer link is free to swivel with respect to the inner link. The rollers are loosely mounted on the bushes so that they rotate when they are engaged with the teeth of the sprocket wheels. This results in rolling friction between the roller and sprocket teeth, reduces friction and results in less wear on them. The pitch of the chain ‘p’ is measured between the axes of adjacent rollers. The width of the chain (b1) is defined as the space between the two inner link plates along the axis of the pin.     CLASSIFICATION OF CHAINS Chains are classified as roller chains and silent chains (inverted tooth or side guide chains). Single roller chain (or simple chain) drives are shown in Fig. 3.3.1. Figure 3.3.2 shows both the single and double roller (duplex) chains. The construction of single roller chain is already explained in the earlier section. A duplex roller chain can be visualized as having two single roller chains placed side by side mounted on same set of pins. The silent chains are heavier, more difficult to manufacture and expensive compared to roller chains, hence their applications are limited. Fig. 3.3.1 Single roller or simple chain on bale opener Fig. 3.3.2 Simple and duplex roller chains on opening and cleaning machine           LUBRICATION OF ROLLER CHAINS   Roller chains must be lubricated to achieve long and trouble-free life. A drop-feed lubrication or lubricant in a shallow bath can be used. A medium or light mineral oil, without additives can be used as lubricants. Heavy duty oils and greases which are highly viscous do not enter the small spaces in the chain parts; and hence they are generally not recommended.     CHAIN TENSION AND BENDING FORCE ON SHAFT   Fig. 3.5.1 shows a pair of chain sprockets transmitting power. The upper part of the chain is in tension and produces the torque on either sprocket. The lower part of the chain is slack and exerts no force on either sprocket. Therefore the total bending force on the shaft carrying the sprocket is equal to the tension on the tight side of chain. Fig. 3.5.1 Forces acting on chain sprockets       GEOMETRICAL RELATIONSHIPS IN CHAIN DRIVE Fig. 3.6.1 shows a sprocket rotating in counterclockwise direction drives a chain. The symbols p, γ , d and z denote for the pitch, pitch angle and pitch diameter and number of teeth on the sprocket respectively. Fig. 3.6.1 Geometry of a chain drive   POLYGONAL EFFECT IN CHAIN DRIVES The number of teeth on the sprocket affects the velocity ratio during its rotation through the pitch angle. The chain passes around the sprocket as a series of chordal-links. This action is similar to a non-slipping belt wrapped around a rotating polygon. For the sack of clarity, a chain drive with sprocket having four teeth is shown in Fig. 3.7.1. Fig. 3.7.1 Polygonal action of chain It is evident that linear speed of the chain is not uniform but varies during every cycle of tooth engagement. This results in a pulsating and jerky motion to the chain and the driven sprocket. This is called “Chordal speed variation”, and is plotted in Fig. 3.7.2.   Animation 3.7.2 Periodic speed variation of chain Fig. 3.7.2 Chordal-speed variation in chain drives For smooth operation of chain drives at moderate to high speeds, it is considered a good practice to use a driving sprocket with at least 17 teeth. Of course, higher number of teeth 19 or 21 gives better life expectancy to the chain with less noise during operation. It is preferable to use an odd number of teeth for the driving sprocket in combination with an even number of chain links for uniform wear and tear on the teeth and rollers. In this case, a particular tooth of the sprocket wheel does not come in contact with a particular link of the chain for every rotation. For space consideration or for very low speeds, sprocket with less number of teeth may be used, by sacrificing the life expectancy of the chain. The failure of roller chains is mainly due to wear on the rollers and pins. The chain is elongated due to wear in the joints and as a result the rollers shift with respect to the teeth on the sprocket. When the sprocket is made of large number of teeth (with shorter pitch), even a small elongation of the chain displaces the chain with respect to the teeth on sprocket to a greater extent, leading to improper meshing. This limits the maximum number of teeth on the sprocket. For roller chains, the recommended maximum number of teeth for the sprocket wheel is about 100 to 120.     APPLICATIONS OF ROLLER CHAINS Few applications of roller chains are listed below: Pedal roller of scutcher. Drives from beater to plain and perforated drums and feed rollers on fine cleaner are through duplex roller chains. Drive from inclined lattice to feed apron and creel apron of bale opener through clutch. Motor to feed-roller, lap winding-roller, and tuft-feeder in high production carding machine. Duplex roller chain transmits motion from main shaft to lap rollers via bottom calender roller on sliver lap machine. Drive to shafts driving the flyers and bobbins on conventional roving machines. Drive to ring rail on ring spinning machines. Drive to creel rollers on drawing machines, and hank meters ( Fig. 3.9.1).   Fig. 3.9.1 Chain drive for hank meter on drawing machine Drive to brush roller shaft on comber. Drive to drafting rollers that feed sheath fibres on friction spinning machine. A drive to creel or table rollers (C, E, G and H) of conventional drawing machine is shown in Fig. 3.9.2. Initially, the drive originates from a driving sprocket, ‘A’ (behind the back drafting roller) to a sprocket, ‘B’ mounted on the shaft of first table roller through a tension sprocket wheel. The sprocket C compounded with the sprocket B drives the sprocket D mounted on the second table roller through a tension sprocket wheel (not shown in the figure) and so on to other table rollers. In high speed drawing machines, chains have been replaced by timing belts. Fig. 3.9.2 Drive to table rollers on a conventional drawing machine                                                                                       X  .  IIII SPUR GEARS  Introduction Gears have specially constructed toothed profile and are extensively used to transmit power in machines. Gears can be classified into spur gears, helical gears, bevel and worm gears. Within these gears there are sub-classification based on designs. Gears are made of ferrous (steel, cast iron), non-ferrous metals (bronze based) and non-metallic materials (Nylon, fibre reinforced in phenolic resin etc.). Steel is the most widely used material for gears. Spur gears are the simplest gears, having the maximum precision and high power transmission efficiency compared to any other gears. Hence, they are preferred as the first choice in industrial machines, except high speed and high load applications. In spur gears, two meshing gears are mounted on parallel shafts. The teeth are cut parallel to the axis of gear. In a normal or external spur gear, the teeth are cut on the outside of the rim of gear ( Fig. 4.1.1). Fig. 4.1.1 Normal or external spur gears on ring spinning machine Generally, the input gear is smaller in size and the output gear is larger in size to get speed reduction. The driver and the driven gears are called ‘pinion’, and ‘gear’, respectively.      DESIGN ASPECTS OF SPUR GEAR Nomenclature of spur gear : The nomenclature of spur gear is illustrated in Fig.4.2.1. The pitch circle shown in the figure will not be visible in an actual gear; but the entire design of gear is based on the pitch circle diameter or the pitch diameter. The pitch circles of a pair of meshing gears must be tangent to each other. The circular pitch, p corresponds to the distance, measured on the pitch circle, from a point on one tooth to a corresponding point on an adjacent tooth. In other words, the circular pitch is equal to the sum of the thickness of a tooth and the space between two adjacent teeth measured along the pitch circle. Fig. 4.2.1 Terminology of spur gears The following notations are used in spur gears: P = Diametral pitch z = Number of teeth d = Pitch circle diameter m = Module (mm) p = Circular pitch (mm) a = Addendum b = Dedendum c = Clearance = b - a do = Diameter of addendum circle or Outside diameter =(d + a) The module ‘m’ is the ratio of the pitch diameter to the number of teeth on the gear. The unit of module in SI system is mm. The diametral pitch P is the ratio of the number of teeth on the gear to its pitch circle diameter. It is the reciprocal of module. The diametral pitch is usually expressed as ‘teeth per inch’. The addendum circle (visible on a gear) is the largest circle on the gear. The addendum ‘a’ is the radial distance between the pitch circle and the addendum circle. The dedendum circle (visible on a gear) is usually the smallest circle on the gear. The dedendum ‘b’ is the radial distance between the pitch circle and the dedendum circle. The dedendum is larger than the addendum; i.e., b > a. The depth of tooth ‘h’ is the sum of the addendum and dedendum; i.e., h = (a + b). The base circle or clearance circle of a gear is tangent to the addendum circle of its meshing gear. The clearance ‘c’ is the difference between the dedendum (b) and addendum (a) of the gears. Geometrical relationships in spur gears :       CONJUGATE ACTION The gears must be designed such that the ratio of rotational speeds of driven and driver gear is always constant. When the tooth profiles of two meshing gears produce a constant angular velocity during meshing, they are said to be executing conjugate action. That is (ω1 / ω2 ) = constant. ................................................................(4.5) Where ω1 = Angular velocity of the driver. ω2 = Angular velocity of the driven. Gears are mostly designed to produce conjugate action. Theoretically, it is possible to select an arbitrary profile for one tooth and then to find a profile for the meshing tooth, which will give conjugate action. One of these solutions is involute profile. The involute profile is universally used for constructing gear teeth with few exceptions. Figure 4.3.1 illustrates a conjugate action.   Animation 4.3.1 Illustration of conjugate action Fig. 4.3.1 Principles of conjugate action       INVOLUTE PROFILE OF GEAR TEETH An example involute profile of gear teeth is shown in Fig. 4.5.1. Two gear-blanks A and B are centered about O1 and O2 respectively. Let us imagine that a cord ‘xy’ is attached to both the base circles of these gears. When the base circles are rotated in different directions keeping the cord always under tension, a point on the cord will trace out two involute profiles, ‘cd’ on the base circle of gear ‘A’ and ‘ef’ on the base circle of gear ‘B’. The involutes are thus generated simultaneously by tracing the points. Fig. 4.5.1 Principles of generation of involute profile for gear teeth Each tracing point on the involutes represents the point of contact between the involutes. The portion of cord ‘ab’ is the generating line. The point of contact moves along the generating line. The generating line is a fixed line, as it is always tangent to the base circles which are fixed. It is clear that the generating line is always normal to the involutes at the point of contact. Thus, the requirement for constant angular velocity ratio is satisfied.    CONSTRUCTION OF INVOLUTE GEAR TOOTH The procedure to construct a gear tooth having an involute profile is given below. Construct a segment of base circle of the gear (centered at ‘O’) . Divide the segment of the base circle into a number of equal parts separated by very small angle ( q ). Construct radial lines OA4, OA3, OA2 etc as shown in Fig. 4.6.1. Click on Image to run the animation Animation 4.6.1 Illustration on construction of involute gear tooth Fig. 4.6.1 Construction of involute gear tooth Construct tangents to the circle A4B4, A3B3, A2B2, A1B1 and from A4, A3, A2 and A1 respectively. Then along the line A1B1, mark the arc-distance A1A0 from A1; along the line A2B2, mark the twice the arc-distance A1A0 from A2 and so on; to get the points B1’, B2’, B3’ and B4’ etc. The curve joining these points starting from A0, would an involute curve A0B1’B2’B3’B4’. In a gear, the involute profile of the tooth starts from the base circle and continues up to the addendum circle with continuously increasing radius. In the procedure described above, the points A0,B1’, B2’, B3’, B4’ are discrete; whereas in actual gear teeth profiling, the θ is infinitesimal that continues points can be generated to get a perfect involute profile for the gear teeth                                                                                     X  .  IIIII CONE PULLEYS ON ROVING MACHINE It has been discussed in module 5 that the bobbin speed is a function of bobbin diameter. The reduction in bobbin speed is effected through cone pulleys (Fig.7.1). After the completion of winding of each layer of roving, the builder mechanism shifts the belt on the cone pulleys, in such a way that the speed of bottom cone pulley is reduced. From the bottom cone pulley, through gear train, drive is transmitted to the epicycle gear train. This speed is superimposed on the fixed speed of the main shaft that corresponds to flyer speed. The output from the epicycle gearing is transmitted to the bobbins through gears. Fig. 7.1.1 Cone pulleys used on a roving machine        DESIGN ASPECTS OF CONE PULLEYS Design aspects of cone pulleys Certain assumptions must be made in order to simplify the approach in designing cone pulleys. The boundary conditions for the design can be set based on the space consideration of whole machine and dimensions related to various elements including gear trains. Based on the space consideration and simplicity of design, the dimensions of the following can be selected. The maximum and minimum diameter for the cone pulleys. Sum of top and bottom cone pulley diameters for any cone belt position is constant as the belt moving over the pulleys is fixed in length. Length of cone pulleys. Based on the dimensions of the following elements, a design criterion must be set. Width of cone belt. Minimum and maximum bobbin diameters of empty & full bobbin to be used. The consideration of the following points could simplify the design approach further. (a) Winding principle employed in winding the roving on to the bobbins is a ‘Bobbin lead’, which is universally adapted on all commercial cotton machines for various practical considerations. (b) The length of the roving delivered by the front drafting rollers per unit time is constant (delivery rate). (c) Twist density along the length of roving is constant (i.e. flyer speed is constant). (d) Roving diameter is assumed to be infinitesimal for very accurate design of the cone pulleys. (e) Roving is incompressible. There is no flattening effect of roving due to winding tension. Roving remains circular even after winding several layers on to the bobbin. Of course, in practice this does not happen. Ideal and accurate design approach should consider the degree of flattening of roving as winding of each layer proceeds. This involves the knowledge of shape, specific gravity and compression resilience of fibres, dimensional changes of certain fibres to moisture absorption, blend compositions, roving tension and twist. However, this is not a serious issue in designing the cone pulleys, once the roving is assumed to be infinitesimal. But these factors do affect the belt initial position and its subsequent shifting. The delivery rate of roving is constant with a well maintained drafting system (i.e., no front roller nip-movement) and the machine is under minimal vibrations (speed variation of front roller due to vibrations are negligible). With perfect gearing and balancing of the flyer mass, the flyer speed is constant. Assumption (d) is useful for minimizing the errors in developing smooth profiles for the cone surfaces. In practical situations, the roving is highly compressible and assumes the shape of an ellipse, with the major axis of ellipse lying along the circumferential direction; and the minor axis of ellipse along the radial direction of the bobbin. In addition, as the roving layers are wound, the radial pressure on the inner layers built up continuously, hence the inner layers deviate further from their circular cross-section, the major diameter of the ellipse increases and the minor diameter decreases till they become incompressible. Further, the rate at which they are compressed with respect to the number of layers are not linear, initially the rate of compression is maximum and may reach zero after winding a certain number of layers. It is very difficult to simulate the conditions of winding of roving, and measure their diameter, especially the minor diameter. Considering the difficulty, the roving diameter may be assumed to be circular and infinitesimal.     STEPS TO DESIGN CONE PULLEYS Steps to design cone pulleys Assume a roving of infinitesimal thickness. Calculate the bobbin diameter (dBm) corresponding to winding of various layers of roving using the Eq. (6.22) Using the Eq. (6.21), (Nw = ν/π dBm), calculate the winding rpm (NW) for winding each layer of roving of infinitesimal diameter. Using the Eq. (6.19), (NW = NB - NF), the bobbin rpm (NB) could be determined. The bottom cone pulley rpm (N2m) for winding the mth layer of roving is determined using the Eq. (6.42) as   The rpm of top cone pulley is 622. The rpm of top and bottom cone pulleys (assuming v as 20.5 m/min. and dB0 < 48 mm) are plotted in Fig. 7.3.1 Fig. 7.3.1 Rotational speed of cone pulleys in relation to number of layers wound The ratio between top and bottom cone diameters (or radii) for winding various numbers of layers, assuming there is no slippage at the interface between the surfaces of cone pulleys and belt is The values of k m are plotted in Fig.7.3.2. (Note: Plotting this ratio is not essential, and only the values are needed for design). Fig. 7.3.2 Ratio of top and bottom cone pulley diameters (km) in relation to the number of layers The radii of cone pulleys, r1m and r2m for various bobbin diameters are related by the expression. r1m + r2m = k0 ............................................................................................................(7.4) Where, k0 is constant by design, as the length of belt moving over the pulleys is fixed. The value of k0 is set at 160 cm.      STEPS TO DESIGN CONE PULLEYS Steps to design cone pulleys Assume a roving of infinitesimal thickness. Calculate the bobbin diameter (dBm) corresponding to winding of various layers of roving using the Eq. (6.22) Using the Eq. (6.21), (Nw = ν/π dBm), calculate the winding rpm (NW) for winding each layer of roving of infinitesimal diameter. Using the Eq. (6.19), (NW = NB - NF), the bobbin rpm (NB) could be determined. The bottom cone pulley rpm (N2m) for winding the mth layer of roving is determined using the Eq. (6.42) as   The rpm of top cone pulley is 622. The rpm of top and bottom cone pulleys (assuming v as 20.5 m/min. and dB0 < 48 mm) are plotted in Fig. 7.3.1 Fig. 7.3.1 Rotational speed of cone pulleys in relation to number of layers wound The ratio between top and bottom cone diameters (or radii) for winding various numbers of layers, assuming there is no slippage at the interface between the surfaces of cone pulleys and belt is The values of k m are plotted in Fig.7.3.2. (Note: Plotting this ratio is not essential, and only the values are needed for design). Fig. 7.3.2 Ratio of top and bottom cone pulley diameters (km) in relation to the number of layers The radii of cone pulleys, r1m and r2m for various bobbin diameters are related by the expression. r1m + r2m = k0 ............................................................................................................(7.4) Where, k0 is constant by design, as the length of belt moving over the pulleys is fixed. The value of k0 is set at 160 cm.    HYPERBOLIC CONE PULLEYS : Combining the Eq. (6.3) and (6.4), the radii of bottom cone pulleys with respect to numbers of layers wound can be found as Substituting the value of r2m in Eq. (7.4), the corresponding values for r1m could be obtained. Mark the r1m and r2m values (i.e., radii of cones) as the radial distances from the axes of both cones, by shifting along the cone axes, each shift equal to l/m (l, denotes length of cone). The initial radii of the cones r11 and r21 can be taken as rmax and rmin respectively. These values definitely correspond to a bobbin of diameter < 48mm. This means that belt is not placed on the edge of cones during winding of roving on bare bobbin whose diameter is 48 mm; in other words, the pulleys are designed to accommodate bare bobbins whose diameters are less than 48 mm to take care of dimensional variation in moulding the bobbins.   STRAIGHT CONE PULLEYS In the case of straight cone pulleys, rmax and rmin are placed apart at a distance equal to the cone length (l) and straight lines connect them ( Fig.7.5.1). Hyperbolic cone pulleys are difficult to design. Further, the belt is always moved on the surfaces of varying inclination. During initial shifting of belt (first few layers of winding), the inclination of cone surfaces is much sharper; and the belt envelops the larger diameters of pulleys and hence, may result in lower winding rate. As a result, cone pulleys today are mostly made straight sided. Fig. 7.5.1 Profile of top and bottom cone pulleys: Straight cones Since the cone length (l) is fixed, Eq. (7.4) also applies to straight cone pulleys. The maximum and minimum radii (rmax and rmin) are fixed and are the same for both top and bottom pulleys. From these, the profiles of cone pulleys are determined. The taper angle of the cone is related to maximum and minimum radii and the length of cone as, tan Υ = (rmax - rmin) / l ......................................................................................(7.6) The value of tanΥ is 0.09123; that is the slope/taper angle is about 5.20. Shifting of belt on straight cone pulleys The radii of top and bottom cone pulleys (r1m and r2m) must vary from their initial values after completion of each layer of winding. The total shift of belt (lm) from its initial position along the axes of pulleys while winding the second layer (of arbitrary thickness) onwards (m = 2, 3, 4…) is related: r1m = r11 - lm tan y ............................................................................................(7.7) r2m = lm tan y ...................................................................................................(7.8) Where, r11 and r21 are the radii of the top and bottom pulleys, while winding the first layer of roving. These are decided as per the values of km obtained from the Eq.(7.9) r11 / r22 = N21 / N1 = km ....................................................................................(7.9) Where, N21 is the rpm of bottom cone pulley while winding the first layer of roving on the bare bobbin, which could be obtained from Eq. (6.42). Note that the rotational speed of top cone is fixed for the given gearing plan. Adding the Eq. (7.7) and (7.8), and substituting k0 = 160, we get r2m + 1m = k0 = 160 ........................................................................................(7.10) Subtracting the Eq. (6.7) from (6.8), we get If the initial values of r11 and r21 are 113.3 mm and 46.7 mm (i.e. rmax for top cone and rmin for bottom cone), then, lm for winding second layer (of arbitrary thickness) onwards is The cumulative axial shifts of belt for both the cases of hyperbolic- and straight-cone pulleys from the extreme left ends of cones are plotted in Fig.7.5.2. The shifting of belt is uniform for the former. For the later, the shifts are longer at the beginning of winding, and then it progressively reduces at later stages of winding. Fig. 7.5.2 Profiles of shifting of belt along the axes of straight and hyperbolic cone pulleys from extreme left end of cones Belt slippage and corrections for belt position. In case of belt slipping over the pulley (usual case), the Eq.(7.2) must be modified taking into account the amount of slippage. If the belt slip by s % over the cone pulleys, then, The belt must be moved towards the starting side of the cone pulleys by some distance to get the right ratio of top- and bottom cone pulley diameters. Considering the belt slipping about 3% (quite normal), the radii of bottom and top pulleys could be set at 106 mm (r11) and 54 mm (r21) respectively while winding the first layer of roving on a bare bobbin diameter of 48 mm. Then, The shifting of belt (that slips by 3%) from the initial position (i.e., r11 and r21 are 106 and 54 mm respectively) is plotted in Fig. 7.5.3. Fig. 7.5.3 Profile of shifting of belt (with 3% slip) on straight cone pulleys from its starting position corresponding to winding on empty bobbin of 48 mm (Total shift is 570 mm) Belt shifting mechanism for straight cone pulleys The unequal but progressively reducing in the shift of belt is achieved by rotating a cam by a fixed degree (that depends on the thickness of roving) is shown in figures 7.5.4 and 7.5.5. Fig. 7.5.4 Cam to control belt shifting on cone Fig. 7.5.5 Belt shifting mechanism for straight cone pulleys One end of a first steel cord is fastened to the cam and wrapped over the cam surface. The other end passes over a fixed guide, roller of belt guide and then attached to the frame of the belt guide. A second steel cord attached to the belt guide passes over a pulley of a stationery frame and weight is hung on to the free end of cord. The roller of belt guide is positioned over a compensating rail that is usually placed parallel to the bottom side of the top cone and the top side of bottom cone. This means that the slopes (taper) of both the cones and compensating rail are the same. The hanging weight always tries to shift the belt guide towards the right side of pulleys (towards the minimum diameter of top cone pulley), but resisted by the cam when it is not in motion. After completion of each layer of winding, a ratchet is rotated by half a tooth. From this, motion is transmitted through gear train to the cam. The angular drift of cam can be changed by the numbers of teeth selected for the traverse change gear of gear train that depends on the coarseness of roving. When the cam turns, the roller of belt guide, rolls over the compensating rail towards right due to the force exerted by the hanging weight, thus, shifting the belt. If the belt shifts by a distance smduring winding of mth layer (which is lm –lm-1), the distance moved by the roller of belt guide over the compensating rail is During first layer of winding, the diameter of the cam is at its maximum, and decreases progressively towards the end. If the working length of cone is 570 mm, the cam can be designed to have 570 radial lengths (R1, R 2, R3 Rm) corresponding to 5700 rotation in order to have fine adjustment of belt shift. The path length of cord between the guide and cam (sm'') progressively increases as the winding proceeds from the first to last layer (Fig. 7.5.6). Fig. 7.5.6 Belt positions, radial position of cam and path length of cords     Further, the circumferential length of cam from R1 to R2 and R2 to R3… Rm-1 to Rm denoted as sm''' progressively decreases and hence, the shifting of belt reduces from second to last layer of winding, evident from the equation. Finer adjustment on belt shifting In spite of precise design and manufacture of cam, belt guide and compensating rails and their controlled operations in moving the belt over the cone pulleys, the roving might encounter slackness or stretch. Changes in humidity, roving twist, and fibre properties such as bulkiness, compressibility, micronair/denier and shape of fibres, affect the effective thickness of roving. Accordingly, the effective bobbin diameter varies, warrants correction on the bobbin speed by changing the incremental movement of belt. The compensating rail comprises of three parts that corresponds to first, second and third quarters of winding, each of its inclination can be changed independently by screws and pivot arrangement. If the roving experiences slackness during first quarter of winding, it implies that bobbin speed is less, which requires slower rates of belt shifting. Then the taper angle (Υ) of the first part of compensating rail (nearest to the cam) must be increased. This reduces the amount of belt shift for each angular drift of cam. When a very high roving tension is encountered during winding, the taper angle must be reduced. Similarly, the remaining parts of winding can be controlled with the help of second and third parts of compensating rail. This requires fair degree experience and skill on the part of operator.       CONE PULLEYS FOR PIANO-FEED REGULATION : Piano feed regulation controls the thickness of lap delivered by the scutcher. Several pedals below the feed roller of scutcher moves up and down independently, depending on the localized variation in the thickness of material. Through links and levers, these movements are mechanically integrated. The integrated output is used to move the belt on cone pulleys, to vary the rotational speed of top cone pulley (output cone pulley). The output from the top cone pulley is transmitted to the feed roller, through gears, thus, adjusting its speed corresponding to average thickness of material sensed at the interface of pedals and feed roller. When the average thickness of material is high, the speed of the feed roller must be reduced and vice versa. The following assumptions are useful in the design of the cone pulleys : The density of web is constant, i.e. mass flow rate of web (g/min.) is proportional to the thickness of web/lap. The sum of the top and bottom cone pulley diameters for any cone belt position shall be constant. Specification for cone dimensions, viz., maximum and minimum diameters, and length. The rotational speed of driving cone (bottom cone) is constant for a given production rate. The mass of a unit length of a web/lap (m) is m = ραt .....................................................................................(7.18) Where ρα is the areal density of web in g/cm2, which is constant; and t is the web thickness in cm. Therefore, the mass flow rate of web (M) in g/min is M = πdf nf ραt .............................................................................(7.19) Where df is the diameter of feed roller in cm; and nf is the rpm of feed roller. When the gear transmission ratio from top cone to feed roller is ‘ e ’; and the rpm of top cone pulley is n2, then, M = πd fe n2 ραt ..........................................................................(7.20) Since, the parameters M, ρα, π , df, and e are constant, under practical condition, then, n2 is proportional to (1/t). Then, 1/t = C1 n2 .................................................................................(7.21) Where, C1 is a constant equal to π df e ρα / M. Also, n2 = (d1n1)/d2 .............................................................................(7.22) Where d1 = Bottom cone pulley diameter in cm   d2 = Top cone pulley diameter in cm   n1 = bottom cone pulley rpm Combining the Eq. (6.21) & (6.22), we get 1/t = C1 n1 (d1/d1) ......................................................................(7.23) Since n1, the rpm of driving cone pulley is kept constant. Therefore, 1/t = C(d1/d1) ............................................................................(7.24) Where, C is a constant equal to C1 n1. By assuming various values for the thickness of web passing between feed roller and the pedals, the ratio d1 /d1 could be found. From this ratio, d1 and d1 values can be computed considering that d1 + d1 = X, a constant. This is given in the Table 7.6.1. Table 7.6.1: The diameters of cone pulleys in piano-feed regulation Web thickness, t (cm) (1/t) = (C.d1/d2) d1+d2 d1 d2 0.1 1/0.1 X calculate calculate 0.2 1/0.2 X     0.3 1/0.3 X     0.4 1.0.4 X     0.5 1/0.5 X     and so on   X     Construct the cones similar to the procedure followed in the design of cone pulleys for roving machine. In practice, the web thickness variation should be considered as infinitesimal to get smooth profile for the cone surfaces. The resulting profiles of the cones are hyperbolic. Straight cone pulleys can also be designed with belt shifting mechanism using a cam and other elements as discussed earlier. When processing different fibres (synthetics and cotton), the parameter, ραmust be considered in the equations (must be kept as a variable; not a part of constant C1), as the fibre specific gravity affects this value.         STRESS-STRAIN RELATIONSHIPS OF MATERIALS Tensile stresses When a component of a material of length (l in mm) is subjected to a static tensile load (F in N) over a cross-sectional area (A in mm2) results in extension (δ, in mm), the tensile stress (δt ) and strain are given by Shear stresses Riveted plates are shown in Fig.8.4.1. When they are subjected to two equal but opposing external forces (F) which are not collinear, the adjacent planes may slide with respect to each other. The stresses on these planes are called direct shear stresses. Fig. 8.4.1 Shear stresses in a riveted joint: (a) Riveted joint (b) Shear deformation (c) Shear stresses The average shear stress on the rivet having cross-sectional area (A) is given by A rectangular element of a component subjected to shear force is shown in Fig. 8.4.2. Fig. 8.4.2 Rectangular element subjected to pure shear force: (a) Pure shear stress; (b) Shear strain Bending stresses A circular shaft which is subjected to a bending moment (Mb) is shown in Fig. 8.4.3. The shaft is subjected to tensile stresses below and compressive stresses above its neutral axis. Fig. 8.4.3 Shaft subjected to bending moment: (a) distribution of bending stresses at the plane X-X; (b) Section of the shaft at section X-X The bending stress at a distance y from the neutral axis or shaft axis is given by Torsional stresses A transmission shaft, subjected to an external torque, is shown in Fig.8.4.4. The torque induces internal stresses in the shaft which resist the action of twist. The internal stresses are called torsional shear stresses. Fig. 8.4.4 Stresses on a shaft due to torsional moment: (a) shaft is twisted (b) distribution of torsional shear stresses The torsional shear stress is give by      FACTOR OF SAFTEY AND ALLOWABLE STRESS In designing a shaft, it is essential to guarantee sufficient reserve strength left on it in case of an accident. It must be assumed that the shaft would be subjected to extremely high load under unforeseeable situations while it is performing. Taking a suitable factor of safety (fs) can ensure this. The factor of safety is defined as The allowable stress is the one, which is used in design calculations to determine the dimensions of shaft. It is considered as a stress, which the designer expects will not be exceeded under normal operating conditions. For ductile materials, the failure stress is limited to the yield stress or yield strength (Syt); and the allowable stress (σ) is For brittle materials, the failure stress is limited to the ultimate tensile stress (Sut); and the allowable stress is   DESIGN OF SHAFT FOR TORSIONAL RIGIDITY Machine tool spindles and some line shafts are designed on the basis of torsional rigidity considerations. The total angle of twist θ in degrees is given by the Eq. (8.20). The permissible angle of twist or limiting value of twist for line shaft and machine tool (or spindle) applications are 3º and 0.25° per m length of shaft respectively. Let us assume that the modulus of rigidity (G) of the shaft material is 79300 N/mm2, and the permissible angle of twist (θ) is 3° /m length of shaft. The distance between the bearings (l) is 600mm. This is the span length of shaft on which the maximum bending takes place and as a result, the maximum deflection also occurs. BRAKES    INTRODUCTION Brake is a machine element which is used either to stop the machine or retard the motion of a moving system, such as a rotating rollers or drums or vehicle where the driving force has ceased to act or is still acting. In practice most brakes act upon drums mounted on the driving shafts or driven shafts. In such cases brake will act either upon the internal surface or external surface of the drum. The brakes acting on the brake drums do not make contact along the whole periphery and the part making contact with the drum is called shoe. The shoe has to expand for internal contact and close in for external contact. When the braking action takes place, the energy absorbed by the brake shoe is converted into heat energy and dissipated to surroundings. Heat dissipation is a serious problem in brake applications     BLOCK BRAKE WITH SHORT SHOE A block brake consists of a rotating drum (brake drum) against which a brake shoe is pressed by means of a pivoted lever (shown in Fig.10.3.1). The block brake is also referred as ‘Drum brake’. Fig. 10.3.1 Drum or Simple block brake The friction force between the shoe and the brake drum acts against the direction of rotation of the drum at the contact region. This causes retardation of the drum. When the friction force is very high, the drum stops rotating. The angle of contact, ‘θ’, between the shoe and the brake drum is usually kept less than 450, to obtain uniform pressure between them. The main disadvantage of the drum brake is the tendency of the drum shaft to bend under the action of normal force, N ( Fig. 10.3.2) Analysis of forces acting on the drum The free body diagram of forces acting on the drum and the lever is shown in Fig.10.3.2. Fig. 10.3.2 Free body diagram of forces acting on the drum and the lever of simple drum brake When the shoe is rigidly attached to the lever, then the torque acting on the brake drum is Mt = μNr............................................................................................................ (10.3) Where μ = Coefficient of friction between the drum and shoe N = normal force acting on the drum (N). The normal force acting on the drum; N = pl1w............................................................................................................ (10.4) Where p = permissible pressure between the shoe and the brake drum (N/mm2) l1 = length of the shoe (mm) w = width of the shoe (mm) The reaction forces on the pivot of the lever in horizontal and vertical directions are denoted as Rx and Ryrespectively. Considering the equilibrium of forces on the lever in horizontal and vertical directions, Rx = μN ..............................................................................................................(10.5) Ry = (N-F) ..........................................................................................................(10.6) Where F is the actuating force on the lever. Taking moment of forces acting on the lever about the pivot point, it can proved that, Depending upon the magnitude of coefficient of friction (μ) and the position of pivot point, there are three possibilities emerge. Case 1 a > μ b .......................................................................................................(10.8) In this case, the friction force (μN ) helps to reduce the magnitude of the actuating force F. From the free body diagram, the moments (Fl), and (μNb ) are counterclockwise. Such a brake is called a partially self-energizing brake. However, the brake is not self-locking, because a small magnitude of positive force F is required for the braking action. Case 2 a = μ b .......................................................................................................(10.9) In this case, the actuating force F is zero, as seen from the Eq. (9.7). This indicates that no external force is required for the braking action. Such a brake is called’ self-energizing’. Since, the actuating force is zero; it is also called as ‘self-locking’ or ‘self acting’ brake. Case 3 a < μ b .......................................................................................................(10.10) Under this condition, the actuating force F becomes negative, as seen from the Eq. (10.7). This results in uncontrolled braking. Drum brakes should be designed so that it is not self-locking and, at the same time, full advantage of the partial self-energizing effect should be taken to reduce the magnitude of the operating force F. 9.3.2 Block brake with short shoe on lap former Block brake with short shoe is used on lap former of conventional blow room. In order to build a compact lap suitable for feeding the card, a brake applies pressure on the lap when it is wound on a spindle. A brake drum is mounted on a shaft which carries two pinions on both sides of the lap former (Fig. 10.3.3). These pinions mesh with racks. Fig. 10.3.3 Brake drum, shoe, racks, pinions and gear train on lap former When winding of lap progress, its diameter increases, pulling the racks up (Fig. 10.3.4). This rotates the pinions. Fig.10.3.4 Racks, pinions and lap spindle A brake shoe mounted on a lever with adjustable weight (W) is in contact with the brake drum during lap winding ( Fig.10.3.5). The upward movement of the lap-rack must overcome the braking action. The work done in raising the racks or lap must be equal to work done in turning the brake drum. Fig.10.3.5 Brake in lap former Balancing the moments about the pivot of the actuating lever, we get the normal reaction force on the lever as, The parameters in the bracket are fixed by design. The hanging weight and its position on the actuating lever are the variable parameters. The force applied on the lap depends on the resilience of fibres under compression. The usual forces applied on the lap while processing synthetic, cotton and viscose rayon are: 40 kN, 25 kN and 18 kN respectively.   DISC CLUTCH AND DICS BRAKE COMBINATION Disk brakes along with disk clutches can be fitted on the main shafts of looms, comber, roving and ring spinning machines and warp beam in warping machines. During emergency situation the machine must be stopped. The connection between the main shaft pulley and main shaft is through multi-disk friction clutch. On the main shaft a multi-disk friction brake can be attached. During emergency situation, the clutch disengages the main shaft from the motor. Then the brake applies on the shaft to stop the machine. While stopping the machine, the disengagement of clutch must precede the braking action by a short time interval, but not in a reverse sequence. Pneumatic actuated multi-disk friction clutch and brake Piston operated multi-disk friction clutch and brake is shown in Fig.10.9.1. A motor transmits motion to the input shaft (main shaft of machine) usually through V-belt. A drum is rigidly mounted on the input shaft which carries a set of disks (A) through bolts. The output shaft is splined and coaxial to the input shaft. Two set of disks (clutch disks B and brake disks C) are placed on the splined sleeve of output shaft, so that they are free to move axially. A stationary outer shaft is coaxial to the output shaft which carries a fourth set of disks (D) through bolts. The arrangement of disks in the clutch is such that the disks A and B are alternately placed with clearance. Similar is the case with disks C and D in the brake. From the output shaft power is transmitted to the machine.   Animation 10.9.1 Operation of multi-disk friction clutch and brake Fig.10.9.1 Multi-disk friction clutch and brake When the motor is stopped, a piston moves towards right. The disks B disengages from the disks A (clutch is disengaged); followed by the disks C contacting the disks D (braking), thus the output shaft and machine will halt rapidly. While starting the machine, the motor first starts rotating; after the motor acquires enough torque, the piston moves from towards left. This disengages the brake (clearance between the disks C and D) first followed by engagement of clutch (contact between the disks A and B); thus transferring power to the machine. Hysteresis disk clutch and brake These types of clutch and brakes provide constant torque for a given control current. They can be used to provide any amount of slip, as long as heat dissipation capacity of the unit is not exceeded. Hysteresis losses transmit torque in this type of clutch and brake. A hysteresis clutch cum brake used on the main shaft of loom is shown in Fig.10.9.2. Click on Image to run the animation Animation 10.9.2 Operation of hysteresis disk clutch and brake Fig. 10.9.2 Hysteresis disk clutch and brake on loom An electro-magnet (clutch magnet) on the input rotor generates magnetic field in the rotor and drag disk. The drag disk is mounted on a splined output shaft and can move along the shaft. A stationary outer shaft is mounted on the output shaft; both are co-axial to each other. An electro-magnet (brake magnet) fixed on the outer shaft can also generate a magnetic field on the drag disk. To transmit motion to the machine, the clutch magnet is actuated which pulls the drag disk away from the de-energized brake magnet. In doing so, the brake is disengaged first followed by engagement of clutch. This transfers drive to the output shaft, and hence the machine. While stopping the machine, motor is turned off. Simultaneously, the clutch magnet is de-energized followed by energizing of the brake magnet. The drag disk moves away from the clutch magnet and towards the brake magnet. Thus, the power transmission is cut off to the machine followed by rapid halting of the machine. The clearance between the drag disk and brake magnet must be lesser compared with that between the drag disk and clutch magnet in order to carry out the operations properly. DISK BRAKES Disk brakes are similar to disk clutches in terms of construction and operating principles. The former is employed to stop or retard the motion of machine or machine elements, whereas the later is for engaging or disengaging a drive system. The force analysis described for the disk clutches (given in module 9) are the same for the disk brakes. Disk brakes are highly suitable for heavy duty industrial application. In yarn winding machine, when yarn breaks or bobbin run-out, braking of the package takes place until the package has come to a complete stop. For this, a brake lining is pressed against brake ring located in the adapter of package. Disk brakes are also used in many textile machines. A disc brake is shown in Fig.10.7.1. Fig.10.7.1 Disk brake            LUBRICATION IN BUSH BEARINGS In a sliding contact bearing, the journal is directly inserted into the bearing. This results in direct metal to metal contact between them. As a consequence the friction is higher between the inner surface of the bearing and the outer surface of journal, if there is no lubricating film present in between them. Bearings can be lubricated with three kinds of lubricants, viz. liquids like mineral oil or vegetable oils, semi-solids like grease, and solids like graphite or molybdenum disulfide. These lubricants are used to reduce friction and wear, dissipate the frictional heat and to protect against corrosion. There are two basic modes of lubrication: (a) thick film and (b) thin film lubrication. Thick film lubrication In thick film lubrication, two surfaces of bearing in relative motion, (viz., the journal and the bearing inner surface) are completely separated by a fluid film. The resistance to relative motion arises from the viscous resistance of the fluid. This does not depend on the structure of journal surface and bearing inner surface as they are not in contact with each other. Thick film lubrication is classified into: hydrodynamic and hydrostatic lubrication. Hydrodynamic lubrication Hydrodynamic lubrication is defined as a system of lubrication in which the load supporting fluid film is created by the shape and relative motion of the sliding elements. The principle of hydrodynamic lubrication in journal bearing is shown in Fig.11.3.1. Fig. 11.3.1 Hydrodynamic Lubrication (a) Journal at rest (b) Journal starts to rotate (c) Journal at full speed When the shaft (centered at O’) is at rest, it goes to the bottom of bearing (centered at O) under the action of load W. This load is due to the weights of shaft and various elements (gears, pulleys) supported by the shaft. The outer surface of journal and inner surface of bearing touch each other during rest, with no clearance at the bottom. The letter ‘e’ denotes the eccentricity, the offset between the axes of the journal and the bearing. As the journal starts to rotate, it will climb the bearing surface. When the speed is increased further, it forces the fluid into the wedge-shaped region between the journal and bearing. As more and more fluid is forced into the wedge shaped region, pressure is generated within the fluid as shown in Fig.11.3.2. This fluid pressure generated in the clearance space supports the external load (W). It can be seen that the pressure distribution around journal varies greatly. Hydrodynamic lubrication does not need a supply of lubricants at high pressure from external source (pumps), as enough fluid pressure is generated within the system. Bearings that use ‘hydrodynamic lubrication’ are called, ‘Hydrodynamic bearings’. Fig. 11.3.2 Pressure distribution in hydrodynamic bearing Hydrostatic lubrication In the case of hydrostatic lubrication, the load supporting fluid film is created by an external source, like pump, supplying fluid at sufficient pressure. Bearing using such system is called ‘externally pressurized bearing’ or ‘Hydrostatic bearing’. The principle of this lubrication is demonstrated in Fig. 11.3.3. Fig. 11.3.3 Hydrostatic lubrication When the shaft is at rest, there is not enough oil pressure as the pump supplying the oil is not yet started ( ( Fig 11.3.3b). When the shaft starts rotating, the pump supply oils at high pressure into the clearance zone. This lifts the journal, reduces the starting friction and also the eccentricity (e). Hydrodynamic and hydrostatic bearings Hydrodynamic bearings are simple in construction, have lower initial and running cost and are easy to maintain compared to the hydrostatic bearings. However, hydrostatic bearings offer the following advantages. High load carrying capacity even at low speeds No starting friction No contact between sliding surfaces at any range of speeds and loads. Thin film lubrication Thin film lubrication is defined as a condition of lubrication where the lubricant film is relatively thin which may rupture that leads to more metal to metal contact; resulting in higher friction. This type of lubrication is also called ‘Boundary Lubrication.’ The boundary lubrication is neither planned by the designer and nor desirable. In sliding contact bearing, heavy load on the bearing, insufficient oil supply, low speed of the journal, chemical reactions and misalignment between the journal and the bearing create conditions for boundary lubrication. Viscosity of lubricants The friction in a lubricated sliding contact bearing is due to the viscous resistance of the fluid in shearing. If a fluid film of thickness ‘t’ and of cross-section ‘A’, is subjected to a shearing force ‘F’ with a relative velocity between the top and bottom surface of the fluid as ‘v’, (refer Fig. 11.3.4) then, Fig. 11.3.4 Fluid film between stationary and moving surfaces         BEARING LOAD AND LIFE       A fan clutch is a thermostatic device that exists as an integral component of certain automotive cooling systems.   Function When the engine is cool or even at normal operating temperature, the fan clutch partially disengages the engine's mechanically driven radiator cooling fan, generally located at the front of the water pump and driven by a belt and pulley connected to the engine's crankshaft. This saves power, since the engine does not have to fully drive the fan. However, if engine temperature rises above the clutch's engagement temperature setting, the fan becomes fully engaged, thus drawing a higher volume of ambient air through the vehicle's radiator, which in turn serves to maintain or lower the engine coolant temperature to an acceptable level. fan clutch for a 1994 lexus LS 400 Mechanical fans are most common in trucks and SUVs, and some RWD cars. This is easier to accomplish because the engine is mounted longitudinally, with the belt accessory components mounted facing the radiator. The fan is mounted on the crankshaft pulley or one of the accessory pulleys (e.g. the water pump pulley) and will spin in between the radiator and the engine, drawing air back through the radiator and blowing it over the engine. Even though the air has been heated by passing through the radiator, it is still much less hot than the engine surface, so the airflow over the engine helps with cooling. In contrast, in a front-wheel drive vehicle, the engine is usually mounted laterally, with the crankshaft and typically all the major accessory shafts parallel to the front axle, so as to directly drive the transaxle; a fan mechanically mounted on an accessory pulley would blow sideways and would not face the radiator. This is why electric engine-cooling fans are used virtually universally in front-wheel drive vehicles. The conversion of mechanical energy to electricity and back to mechanical rotary power with a fan motor is less efficient than a direct mechanical connection, but this is more than compensated by greater control of an electric fan through electronic thermostatic controls which can turn the fan completely off when the engine temperature is below the setpoint. Types Most fan clutches are viscous or "fluid" couplings, combined with a bi-metallic sensory system similar to that in a thermostat. Some clutches are electronically controlled (instead of bi-metallic strip). These provide the potential to control the level of engagement depending on any number of inputs. Common controlling factors might include engine oil temperature, transmission oil temperature, coolant temperature, AC system pressures and ambient air temperature. Failing symptoms A fan clutch is as reliable as any other component on a vehicle, yet sometimes they fail. A common symptom of fan clutch failure is overheating at idle or in heavy traffic. Bad fan clutches can also cause poor performance of the car's air conditioning system because the fan also cools the air conditioner's condenser, which is directly in front of the radiator. A fan clutch can also fail in a stuck on position where engine power is lost even when the fan is not necessary. This type of failure can also decrease fuel economy. Another potential symptom of fan clutch failure (always drawing air at a high rate) in a cold weather climate is that the fan / engine sounds excessively noisy and fuel consumption increases. You may find that the heating system blows lukewarm air never delivering sufficient hot air but as this is also controlled by the cooling system thermostat it's not the best indicator. Another symptom is an extra (abnormal) grinding sound from the engine area, which may be caused by worn bearings in the fan clutch. Automotive industry            Modern assembly line Play media A video showing new SEAT, Škoda & Volkswagen cars being transported by rail at Kutná Hora město train station in the Czech Republic The automotive industry is a wide range of companies and organizations involved in the design, development, manufacturing, marketing, and selling of motor vehicles, some of them are called automakers. It is one of the world's most important economic sectors by revenue. The automotive industry does not include industries dedicated to the maintenance of automobiles following delivery to the end-user, such as automobile repair shops and motor fuel filling stations. The term automotive was created from Greek autos (self), and Latin motivus (of motion) to represent any form of self-powered vehicle. This term was proposed by Elmer Sperry. Contents    a cloud story History of the automobile Thomas B. Jeffery automobile factory in Kenosha, Wisconsin, c.1916 Citroën assembly line in 1918 The automotive industry began in the 1890s with hundreds of manufacturers that pioneered the horseless carriage. For many decades, the United States led the world in total automobile production. In 1929, before the Great Depression, the world had 32,028,500 automobiles in use, and the U.S. automobile industry produced over 90% of them. At that time the U.S. had one car per 4.87 persons.[3] After World War II, the U.S. produced about 75 percent of world's auto production. In 1980, the U.S. was overtaken by Japan and became world's leader again in 1994. In 2006, Japan narrowly passed the U.S. in production and held this rank until 2009, when China took the top spot with 13.8 million units. With 19.3 million units manufactured in 2012, China almost doubled the U.S. production, with 10.3 million units, while Japan was in third place with 9.9 million units.[4] From 1970 (140 models) over 1998 (260 models) to 2012 (684 models), the number of automobile models in the U.S. has grown exponentially.[5] Safety  Automobile safety Safety is a state that implies to be protected from any risk, danger, damage or cause of injury. In the automotive industry, safety means that users, operators or manufacturers do not face any risk or danger coming from the motor vehicle or its spare parts. Safety for the autmobiles themselves, implies that there is no risk of damage. Safety in the automotive industry is particularly important and therefore highly regulated. Automobiles and other motor vehicles have to comply with a certain number of norms and regulations, whether local or international, in order to be accepted on the market. The standard ISO 26262, is considered as one of the best practice framework for achieving automotive functional safety. In case of safety issues, danger, product defect or faulty procedure during the manufacturing of the motor vehicle, the maker can request to return either a batch or the entire production run. This procedure is called product recall. Product recalls happen in every industry and can be production-related or stem from the raw material. Product and operation tests and inspections at different stages of the value chain are made to avoid these product recalls by ensuring end-user security and safety and compliance with the automotive industry requirements. However, the automotive industry is still particularly concerned about product recalls, which cause considerable financial consequences. Economy  Automotive industry by country Around the world, there were about 806 million cars and light trucks on the road in 2007, consuming over 980 billion litres (980,000,000 m3) of gasoline and diesel fuel yearly. The automobile is a primary mode of transportation for many developed economies. The Detroit branch of Boston Consulting Group predicts that, by 2014, one-third of world demand will be in the four BRIC markets (Brazil, Russia, India and China). Meanwhile, in the developed countries, the automotive industry has slowed down. It is also expected that this trend will continue, especially as the younger generations of people (in highly urbanized countries) no longer want to own a car anymore, and prefer other modes of transport. Other potentially powerful automotive markets are Iran and Indonesia. Emerging auto markets already buy more cars than established markets. According to a J.D. Power study, emerging markets accounted for 51 percent of the global light-vehicle sales in 2010. The study, performed in 2010 expected this trend to accelerate. However, more recent reports (2012) confirmed the opposite; namely that the automotive industry was slowing down even in BRIC countries. In the United States, vehicle sales peaked in 2000, at 17.8 million units            Automotive Engineering Basic Structure of an Automobile Automobile Engineering deals with designing, manufacturing and operating automobiles. Automobile basic structure of motorcycles, buses, trucks    Automobile Vs Automotive An automobile is also used as a synonym to automotive. Automobile engineering focuses on the cars. It manages the design, research, manufacture and operation of ground based vehicles. Automotive Engineering particularly deals with the construction and design of automobiles. Automotive is used as a descriptive word that indicates something that is related to or concerned with engine vehicles. In common usage, automotive engineering is same as automobile engineering.     Specializations qualification at automotive area : Electronics and control systems Fluid mechanics Rapid prototyping Computer-based systems and electronics Engine systems Vehicle and pedestrian safety or supply chain management Fuel technology and emissions Emissions control Aerodynamics Thermodynamics “ Some foreign automobile companies like Renault, Volkswagen, Audi etc now focusing on India as a base to manufacture cars.   Toyota