Handbook of Conveyor & Elevator Belting - Goodyear P1

Handbook of Conveyor and Elevator Belting on CD

Metric Edition

The Goodyear Tire & Rubber Co's Handbook of Conveyor and Elevator Belting is the longest running belt conveyor technical publication in the world. First published in 1921 and upgraded numerous times, this world renowned reference has been used by engineers in all fields of bulk material handling across the globe.

Now available in a digital format, it has become even easier to use and an even more valuable reference and design tool. Browse through the contents from the menu to the left, or look for specific information with the search tool in the upper left hand corner.

(C) Copyright 2000, The Goodyear Tire & Rubber Company all rights reserved

Disclaimer & Warranty and Conditions of Sale

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INTRODUCTION

1-1. HISTORY

Transporting bulk materials by conveyor belts dates back to approximately 1795. Most of these early installations handled grain over relatively short distances.

The first conveyor belt systems were very primitive and consisted of leather, canvas, or rubber belt traveling over a flat or troughed wooden bed. This type of system was not an unqualified success but did provide incentive for engineers to consider conveyors as a rapid, economical, and safe method of moving large volumes of bulk materials from one location to another.

During the 1920s, the Colonial Dock installation of the H. C. Frick Company showed what belt conveyors could do in long-distance hauling. This installation was underground and handled run of mine coal over some 8 kilometers. The conveyor belt consisted of multiple plies of cotton duck and natural rubber covers, which were the only materials used to manufacture belting at that time. Although outmoded by today's standards, this material handling system was selected in preference to rail haulage, which has proven to be a proper choice.

During World War II, natural components became so scarce that the rubber industry was forced to create synthetic materials to replace them. Today, The Goodyear Tire & Rubber Company produces belting with an almost endless list of polymers and fabrics to meet the design requirements of any conveying situation. Possible uses of conveyor belting have broadened considerably since the Frick installation.

The basic advantages of conveyors over other means of transport (such as truck, rail, skip-hoist, and aerial tramway) for bulk haulage are numerous. The following paragraphs indicate why today's belt conveyors have become the primary method for bulk material handling.

1-2. CAPACITY

Conveyor belts have no equal in capacity among competing transport means. At a belt speed of 5 m/s, a 1600-mm-wide conveyor belt delivers more than 100 metric tons per minute of material that is 1000-kg/m3 material.

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1-3. ADAPTION TO GROUND PROFILE

Conveyors can follow ordinary natural cross-country terrain by virtue of their ability to traverse relatively steep grades (up to and including 18 degrees, depending on the material being carried). With the development of high-tension synthetic fabrics and/or steel cord reinforcing members, one flight can extend for several miles with both vertical and horizontal curves.

1-4. ROAD BED

A belt conveyor system operates on its own bed of idlers, thus requiring a minimum of attention. Repair or replacement is both fast and easy, and the cost of routine maintenance is minimal.



1-5. LEAST DEGRADATION OF MATERIAL

The smooth ride of conveyor belt systems produces little degradation of the material being conveyed.

1-6. ENVIRONMENTAL CONSIDERATIONS

Electrically powered conveyor belt systems are quiet (an important feature in procuring right-of-ways and in complying with the Occupational Safety and Health Act regulations). Belt systems can be covered to help ensure clean air. They even can be buried out of sight for quiet, functional, and aesthetic reasons.



1-7. MINIMAL LABOR

One worker per 2 kilometers is generally adequate to monitor a properly designed belt conveyor system. Contrast this with the number of drivers on a truck operation handling equal tonnage.

1-8. LIGHT WEIGHT OF CONVEYOR STRUCTURE

Low weight of load and conveyor structure per linear meter allows simple structural design for bridging gullies, streams, highways, or similar obstacles. Likewise, a conveyor structure on a hillside requires little excavation and does not invite hazards from earth or rockslides. Because the structure is compact, it requires a minimum of covering for protection.



1-9. MULTIPLE GATHERING AND DISCHARGE POINT CAPABILITIES

These capabilities are important in mining or excavation where two or more digging operations can feed to a central loading point. At the discharge end, material can be dispersed in several directions from the main line. Material can be also discharged at any point along the conveyor line by using a tripper. Pendulum or caterpillar-mounted belts can be swung in a 180-degree arc to follow a digging shovel or can be used on the discharge end for stockpiling.

1-10. MOBILITY AND EXTENSIBILITY

Modern modular conveyor lines can be extended, shortened, or relocated with a minimum of labor and time.



1-11. LOWEST POWER REQUIREMENT

Conveyors require the least power per ton of any means of haulage. Decline conveyors, depending on degree of slope, often generate power that can be fed back to the line for other uses. The Goodyear ST5800 steel cable belt is the highest strength regenerative belt in North America.

Phelps Dodge - Copper Mine

1-12. CONTROL

Properly designed conveyor systems have control reduced to pushbutton proportions and can be self-controlled by interlocking limit switches.



1-13. EARLY WARNING OF FAILURE

Generally, belts signal their failure from wear many months in advance. With proper safety devices and safeguards, accidental damage can be minimized and contained.

1-14. SAFETY

Bulk material transport by conveyor belting is inherently safer than other methods, particularly in coal mines where safety records show very favorable results with belt systems versus rail haulage.



1-15. ALL-WEATHER CAPABILITIES

For minimal cost, belt conveyors can be protected from rain, snow, and other inclement weather that could adversely affect rail or truck haulage.



FUNDAMENTALS OF BELT DESIGN

2-1. INTRODUCTION

Many engineers and conveyor belt users are familiar with the theory and fundamentals of transmission belting. An analysis of the general aspects of conveyor belt driving relative to transmission belting provides a foundation for the design of belt conveyors and belt elevators. In both conveyor and transmission belting, a belt is driven by friction between the belt and drive pulley or pulleys. Certain other design elements also are much the same whether power is being transmitted or materials transported. The similarity between the two cases can be observed in the following discussion of the fundamentals of belt design as related specifically to belt conveyors and belt elevators.

2-2. DEFINITIONS

Tension in a belt is a force acting along the length of the belt and tending to elongate it. Belt tension is measured in newtons. When the tension is referred to some unit of cross-sectional area, it is known as unit tension and is measured in kilonewtons per metre (kN/m)..

Torque is the effectiveness of a force to produce rotation about an axis and thus involves the size of the force and its moment arm. Torque is the product of a force (or tension) and the length of the arm through which it acts and is expressed in newton-metres (N · m)..

Energy and work are closely related and are expressed in the same units. Work is the product of a force and the distance through which it acts. Energy is the capacity for performing work. Each is expressed in joules, where a joule is one newton-metre. The energy of a moving body, in joules, is given by:

,

where

m = mass in kilograms,

v = velocity in metres per second, and

E = energy in joules.

Power is the rate of doing work or transmitting energy. The mechanical power unit is the watt, which is defined as one newton-metre per second.

Power expended for a period of time produces work, giving rise to the term kilowatt-hour.

2-3. COEFFICIENT OF FRICTION

If, as in Figure 2-1, a body of weight (w) rests on a horizontal plane surface and a force (p) parallel to the surface is just enough to cause the body to be at the point of slipping, the ratio of p to w is the coefficient of friction (f) between these surfaces. Thus,

. (2-1)

Otherwise stated, coefficient of friction is the ratio of tangential to normal force when slip is about to occur.

Figure 2-1 Coefficient of Friction

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2-4. TENSION RELATIONSHIPS

Consider a rope or belt, as in Figure 2-2, hanging over a pulley that resists rotation. Tensions TA and TB are caused by large and small weights, respectively. Common experience teaches that, if the coefficient of friction between belt and pulley is large enough, a considerable difference in tension is possible in such a system. Experience also tells us that, when the arc of contact is reduced (as in Figure 2-3 with a freely turning idler), TB must be larger to keep the belt from slipping. The essential factors are the tensions, the coefficient of friction, and the angle or arc of contact.

If, in Figure 2-2 or 2-3, the unbalanced tension (TA-TB) is large enough to overcome the resistance, the pulley will turn, but the action is limited by the length of the belt. It is an easy step to Figure 2-4, where a joined or endless belt is applied to two pulleys. A turning moment or torque applied at shaft O1 causes a torque at shaft O2. Thus, the action described in Figure 2-2 is applicable continuously in a system like Figure 2-4, thus illustrating the fundamental tension relations in belt driving.

To find the relation of TA coefficient of friction (f), and the arc of contact (a) in radians, refer to Figure 2-5, which represents a very small element of the belt of Figures 2-2, 2-3, or 2-4. The tension in the belt at b is T, and at a it is (T + ∆ T) due to friction. The element "ab" subtends the very small angle ∆ a. The forces are more clearly represented in Figure 2-6, which shows that the force Fn between this portion of the belt and the pulley is given by:

(2-2)

(here ∆ T is negligible).

(the belt being at the point of slipping).

(2-3)

Figure 2-2 Freely Turning Idler Pulley

Figure 2-3

Figure 2-4



Taking the limits as ∆ a approaches zero,

.

Integrating,

,

,

, and (2-4)

, (2-5)

where "a" is the arc of contact in degrees, or

(2-6)

Equations 2-4, 2-5, and 2-6 are valid only under the conditions for which they were derived: (1) the belt is at the point of slipping and (2) centrifugal tension is not included in TA or TB (see also Equation 2-12). Note that in Equations 2-4, 2-5, and 2-6 the left-hand term is the ratio of tensions; therefore, the only requirement for the tension units is that they be the same for both TA and TB.

Figure 2-7 illustrates the manner in which the ratio of tensions can be built up with increasing arc of contact. The larger arcs, of course, require multiple driving pulleys. The vertical scale gives the number of newtons of tight-side tension possible for each newton of slackside tension for f = 0.35.

Figure 2-5 Force Between Belt and Pulley

Figure 2-6 Force Between Belt and Pulley

Figure 2-7 Arc of Contact versus Tension Ratios



2-5. CENTRIFUGAL TENSION

In the usual conveyor and elevator installation, centrifugal tension is a negligible factor because the speeds are relatively slow. For the occasional case where this tension should be considered, a brief discussion is included.

The centrifugal force, Fc (N), acting on a mass of "m" kilograms, moving "v" metres per second in a curved path of radius R metres, is given by:

. (2-7)

If, instead of representing the mass of the body, "m" is kilograms per metre of length, the centrifugal force for the element in Figure 2-5 is given by:

(2-8)

The centrifugal force acting on the elements of a belt is balanced by centrifugal tension (Tc) in the belt. From a relationship similar to Figure 2-6,

. (2-9)

From Equations 2-7 and 2-8,

.

Taking limits as ∆ a approaches zero,

. (2-10)



Now, if

,

. (2-11)

Note that "w" for calculating Tc is in kilograms per metre of length of belt of the same cross-sectional dimensions as those used for T1, T2, TA, TB, and Tc. Thus, if the tensions are expressed in newtons per ply-millimetre, "m" must be given in kilograms per ply-millimetre per metre of length.

Where it is understood Tc is negligible, Equation 2-1 is commonly written:

. (2-12)

In this book, where speeds are known to be relatively low, Equation 2-12 is used.

2-6. CREEP

Belt creep exists whenever a belt passes around a pulley and there is a difference between the entering and leaving tensions. Consider a portion or element of belt approaching a driving pulley. If the tension is high with reference to the torque, the belt element will travel at the same speed as the pulley face through some part of the arc of contact. Through the remainder of the are of contact, this portion of belt will be under progressively less tension down to the slackside tension at the exit point. During the slackening process, the belt element shortens (recovers from elongation) and consequently moves slower than the pulley face. This relative motion is creep.

If the load is increased, the arc in which creep occurs (the arc of creep) increases. If the load is sufficiently increased, the arc of creep may become as large as the arc of contact, in which case the belt will be at the point of slipping. The remedy, of course, is to provide more slackside tension.

Whether the belt is being driven by a pulley or is itself driving a pulley, the arc of creep always starts from the exit point and progresses toward the entry point as the load increases.

Consider the action in the vicinity of the drive of a conveyor or elevator belt. If E is the dynamic modulus of elasticity of the belt, and v1 and v2, are the entry and exit velocities, respectively,

. (2-13)



Since T1 is small compared with E,

(approximately). (2-14)

Consider the belt velocity as it approaches the drive pulley as a base. The belt slows down by the amount of creep percentage where it leaves the drive pulley. The recovery of this velocity loss is spread over the entire conveyor or elevator in a manner depending on the tension cycle.

While the creep percentage is usually small enough to be neglected without appreciable error, it can be significant in cases such as weightometer calculations.

2-7. TORQUE AND POWER FORMULAS

In the following equations:

P = power in kilowatts,

S = belt speed in metres per second,

Te = effective tension in kilonewtons, and

rpm = revolutions per minute.

Thus,

. (2-15)

. (2-16)

. (2-17)

. (2-18)

. (2-19)

2-8. K VALUES

As used in Section 9, K is the ratio of slackside tension to effective tension for the applicable conditions of arc of contact and coefficient of friction. K is used in the determination of the slackside tension required for driving with a given effective tension and specified operating conditions of type of takeup, pulley surface (coefficient of friction), and arc of



contact.

Takeup provisions are included by using a lower coefficient of friction for manual takeup. In this way, a reserve of tension is provided initially to eliminate too frequent adjustment of the manual takeup mechanism.

Since

and

,

, (2-20)

where R = ratio of tensions.

K values from Equation 2-22 for the more commonly used arcs of contact and for both manual and automatic takeups are listed in Table 9-A and Table 9-B. K values calculated from Equation 2-22 for dual or tandem drives are accurate only when there is a proper distribution of power between the primary and secondary drives. Section 9 contains a complete discussion of this and provides a means of determining K for such drives regardless of the power distribution.

2-9. FACE PRESSURE

In large, heavy-duty installations, it is advisable to check the face pressure to make sure that safe limits are not exceeded. The analysis follows Figures 2-5 and 2-6.

Since

Taking limits as approaches zero,

. (2-21)

where

P = face pressure in kilopascals (kPa),

r = radius in metres, and

T = kilonewtons per metre width.



Extreme face pressures in conveyor belt applications are rarely if ever encountered because pulleys are normally large enough for other reasons to be well within safe limits from a pressure standpoint. The problem from extreme pressure would be in the form of peeling or rolling away of a pulley cover, especially at drive pulleys where there also would be some creep.

High-tension steel cord belts are usually the only place where face pressure becomes a factor in pulley selection. The pulley diameter must be selected to a maximum of 1000 kPa.



CONVEYOR BELT CONSTRUCTION

3-1. ELASTOMERS

A. Introduction

The contemporary rubber chemist enjoys a broad spectrum of basic rubbers and compounding materials with which to build better and longer lasting rubber products. Whereas in the past natural rubber played the major role in the construction of conveyor belts, it now shares the lead with many very versatile elastomers.

In the last few decades, manmade rubbers have come into use that offer advantages in economics and properties. The ever-widening application of conveyor belts has demanded and expanded the skills of the rubber chemist and development engineer. Many of these new materials require unusual methods in compounding, processing, and assembling of conveyor belts, with the net result being a more satisfactory product. The new materials enable the conveyor belt manufacturer to accept operating conditions involving swelling oils, corrosive materials, and hot loads, knowing that good service can be provided.

The following paragraphs describe briefly some features of the more popular elastomers used in conveyor belts. Table 3-A summarizes these elastomers and their typical properties.

B. Natural Rubber

The term "natural rubber" refers to the rubber-like materials produced by the coagulation of a plant or tree latex. Of the many, many varieties, the Hevea tree has achieved the greatest commercial use.

Natural rubber and synthetic polyisoprene, Natsyn®, can be compounded to provide good tensile strength and elongation over a wide range of hardnesses. They provide excellent characteristics of resiliency and elasticity. Where these characteristics are important, natural and synthetic "natural" rubber can be expected to give the best results as compared to other elastomers.

Natural rubber has excellent low-temperature resistance. In addition, it can be formulated to provide good abrasion and cut-tear resistance, which are desirable properties in many conveyor belt installations.

C. Plioflex®

Plioflex®, SBR, is made by the copolymerization of styrene and butadiene and falls in the classification of non-oil-resistant synthetic rubber. Good abrasion resistance, low temperature flexibility, and low water absorption are characteristic of this rubber.

Styrene-butadiene rubbers can be polymerized by either a hot or cold process. The "cold rubber" grade exhibits improved abrasion resistance and has attained extensive use in belting.

D. Butyl and Chlorobutyl

Butyl rubber is a copolymer of isobutylene-isoprene and attained initial popularity because of a low rate of gas permeation. Good resistance to weathering, heat, and chemical attack expanded its application to many products. It exhibits some resistance to vegetable oils but is classed as a non-oil-resistant elastomer.

Chlorobutyl is the product of chlorinating butyl rubber, which results in many advantages. This chemical modification enhances the heat resistance of the elastomer, widens compounding boundaries, and improves processing.

Both butyl and chlorobutyl have been used for conveyor belts to utilize their resistance to heat and chemical attack.

E. Ethylene Propylene

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Ethylene propylene diene monomer, EPDM, is recognized for its great resistance to weathering and ozone attack. EPDM can be compounded for exceptional heat resistance as well as resistance to corrosive materials. Good abrasion characteristics are obtainable through proper formulation of this elastomer. The properties offered by EPDM make it an attractive building material for conveyor belts handling chemicals and hot materials.

F. Chemigum®

The Chemigum® rubbers, butadiene-acrylonitrile copolymers, are designed to provide a high degree of oil resistance. These elastomers also can be compounded for good natural and heat aging characteristics and abrasion resistance. They are specialty polymers and are used where their properties are advantageous for specific conveyor belt installations.

G. Chloroprene

There are many types of chloroprene rubber, also known as neoprene, which provide a wide range of properties for product design. All types produce vulcanizates with some resistance to oils, fats, and greases. Chloroprene has ozone and sun-checking resistance as well as good abrasion resistance. Resistance to flame propagation, cut growth, moderate heat, and many chemicals is characteristic of the chloroprene vulcanizates, which makes the elastomer useful for specialized conveyor belting.

H. Hypalon*

Hypalon®, chlorosulfonated polyethylene, is a medium oil-resistant polymer. In addition, Hypalon® provides good resistance to moderate heat, many chemicals, and weathering. Lightly colored compounds having good physical properties maintain their color and durability after long outdoor exposure. Hypalon® does not support combustion. It offers good dynamic properties and resistance to abrasion.

I. Urethane

Urethane is a rather remarkable elastomer with unusually high physical properties. The urethanes are available in both liquid and millable form. The liquid urethanes can be cast or sprayed, whereas the millable type is processed in the same manner as natural rubber. Its exceptional tear and abrasion resistance, coupled with its oil resistance, makes it useful for belts subject to impact and operating in an oil environment.

J. Fluoroelastomers

The fluoroelastomer family is noted for resistance to mineral oils and most solvents at high temperatures. The fluoroelastomer belt cover offers a low coefficient of friction and an excellent non-sticking surface for handling tacky materials. These qualities make it attractive for the unusual belt installation.

*®Registered tradename of E. I. DuPont de Nemours Co., Inc.

Table 3-A Elastomer Characteristics

Commercial Name Type ASTM nomenclature Typical Characteristics

Natural Rubber Isoprene, natural NR

A good balance of high resilience, tensile strength, and tear resistance. Good wear properties, low permanent set, and good flex qualities at low temps. Superior tear and cut-growth resistance.

Natsyn® Isoprene, synthetic IRHigh resilience, tensile, tear and cut-growth resistance. Good low-temp properties.

Plioflex® Styrene and butadiene SBR

Good mechanical properties ranging slightly below those of natural rubber. Can be compounded for good abrasion, wear, and tensile properties.



Butyl and chlorobutyl Isobutylene-isoprene, and butyl, chlorinated IIR, CIIR

Low permeability to gases and vapors. Excellent dampening properties. Resists aging from weather, ozone, heat, and chemicals. Dielectric properties are good.

Ethylene proplylene terpolymer

Ethylene, propylene, and non-conjugated diene EPDM

Excellent ozone, oxygen, and weather resistance. Good color stability, dielectric qualities, and low temperature properties. High elasticity and good heat resistance.

Chemigum® Butadiene and acrylonitrile NBR Excellent resistance to solvents, fats, and oils, and aromatic hydrocarbons. Good aging properties and good abrasion resistance.

Chloroprene Polychloroprene CR

Good oil and chemical resistance. Heat and flame resistant. Good resistance to oxidation, heat, and abrasion.

Hypalon® Chlorosulfonyl-polyethylene CSM

Good ozone resistance and light stability with excellent resistance to weather, heat, and abrasion. Good chemical resistance to many acids and alkalis. Oil and grease resistant.

Urethane Polyester or polyether polyols and di-isocyanates

High abrasion resistance, tear strength, and tensile strength. Good elongation, excellent shock absorption with a wide range of flexibility and elasticity. Good solvent resistance to lubricating oil and fuels.

Flouroelastomers Vinylidene fluoride and heraflouropropylene FPM

High temp resistance with good thermal stability. Resistant to oils, solvents, fuels, and corrosive chemicals.

3-2. REINFORCEMENTS

A. Fibers

A variety of synthetic or manmade fibers are now heavily used in conveyor belts where cotton once was predominant. Vastly increased strength requirements, the desire to obtain equivalent strength at lower cost, and the need for special materials for certain applications all have led to these developments. Table 3-B lists characteristics of some of the more common materials used to make up today's conveyor and elevator belt carcass fabrics.

Table 3-B Belting Fiber Characteristics

Characteristics CottonRayon

(viscose, high tenacity)

Nylon (high tenacity) Steel Glass Polyester Asbestos

Specific gravity 1.55 1.53 1.14 7.8 2.50 1.38 2.6

Tensile (MPa) 410 to 616 405 to 770 604 to 956 2275 2118 731 to 1157 505 ave

Tenacity* (mN/tex)

264.8 to 397.2 264.8 to 503 529.6 to

838.5 --- 847.3 529.6 to 838.5 194.2**

Tenacity, wet (percent of dry) 100 to 130 61 to 75 84 to 90 100 92 70 ---

Elongation at break (percent)

3 to 7 9 to 26 16 to 28 1 to 2 2 to 3 10 to 14 6

Fiber diameter (mm)

0.0178 to 0.0203

0.0102 to 0.0381

0.0076 and up

0.1016 to 0.508

0.0076 to 0.0102 0.0076 and up 0.0000178



*Tenacity is strength per unit of weight as contrasted to the usual structural designation of strength per unit of cross-sectional area. It is expressed in millinewtons per tex, where tex is the mass in grams of 1000 metres of fiber or yarn. Thus, a 94-tex rayon yarn has a mass of 94 grams per 1000 metres of length. Tensile in MPa = tenacity in mN/tex x specific gravity.

** Tenacity, as measured in the yarn, is much lower than the fiber tenacity for cotton and asbestos, but this loss is small for continuous-filament yarn such as rayon nylon, and polyester. Hence, yarn strengths are relatively greater in rayon, nylon, and polyester.

3-2. REINFORCEMENTS

B. Yarns and Weaves

1. General

Fibers are made into yarns which, in turn, are woven into the belt fabrics made of warp yarns (running lengthwise) and filler yarns (running transversely). Some of the more common fabric weaves are as follows:

2. Plain Weave

Most belt carcass fabrics are formed with a plain weave; that is, the warp and fill yarns alternately cross each other as illustrated in Figure 3-1.

Since the warp yarns are the tension or load-carrying members, the fabrics are designed with the dominant strength in this direction. This can be accomplished by using a greater number of ends per centimeter in the warp; by using larger, stronger yarns in the warp; or, in some instances, by using a combination of both of these factors.

Figure 3-1 Plain Weave

3. Filling Rib Weave

This weave is similar to the plain weave except it has double warp yarns and single filler yarns (Figure 3-2).

Figure 3-2 Filling Rib Weave

4. Woven Cord

This cord is shown in Figure 3-3 and consists mainly of warp yarns held in position by widely spaced, very fine fill yarns. There is no crimp (waviness in the yarn) in cord fabrics as in the plain weave so that stretch must be controlled only by the nature of the fiber used and the amount of twist in the yarn.

Figure 3-3 Woven Cord



5. Straight Warp Weave

In this weave, the tension bearing warp yarns are essentially straight with little or no crimp. Fill yarns are laid above and below the warps, and the warps and the fills are held together with binder warp yarns (see Figure 3-4).

Figure 3-4 Straight Warp Weave

6. Solid Weave

This weave consists of multiple layers of warp and fill yarns held together with binder warp yarn (Figure 3-5).

Figure 3-5 Solid Weave

7. Leno Weave

This weave is an open mesh weave commonly used with breaker fabrics (see Figure 3-6).

Figure 3-6 Leno Weave



3-2. REINFORCEMENTS

C. Carcass Fabrics

Years ago, most textile reinforced conveyor belts utilized fabric plies made of cotton, cotton-nylon, rayon, or rayon-nylon. For improved cover adhesion and abuse resistance, a breaker was often placed between the cover and the carcass.

Through the 1960s and 1970s, carcass reinforcements underwent a change until, today, most belts are made with fabrics of nylon, polyester, or combinations of the two. These fabrics are superior to the older fabrics in nearly all respects including strength, adhesion, abuse resistance, fastener holding, and flex life. Breakers are rarely needed or used with these belts because little or no improvement is achieved in either adhesion or abuse resistance.

Aramid fabrics (The Goodyear Tire & Rubber Company Flexten®) are gaining in conveyor belt usage and offer high strength, low elongation, and heat resistance.

3-3. THE ASSEMBLED BELT

A. Carcass Constructions

1. Plied Carcass

Ply belts such as The Goodyear Tire & Rubber Company Plylon Plus® and Glide® can be made in custom widths to meet the strength of a multiple-ply belt and with a much heavier skim coat between plies.

These belts can be made in custom widths with capped edges, but they are normally stocked in long wide rolls from which individual belts are then cut to ordered width and length. With today's fabrics, the cut edges will give service equal to or exceeding that of rubber edges (see Figure 3-7).

Figure 3-7

Plied Carcass Belt (The Goodyear Tire & Rubber Company Plylon Plus®) with Cut Edges

2. Steel Cord

Steel cord belts are supplied in two different constructions, each of which uses a uniplane layer of galvanized or brass-plated steel cords as the tension member. These belts can be built with great strength at little sacrifice in flexibility, and they are virtually free of stretch or shrinkage. Fabric belts of similar strength would be impractical because they would be too thick and stiff either to trough or bend around normal size pulleys.

Figure 3-8 illustrates the all-gum type steel cord belt which, as its name implies, consists only of rubber and steel cord. A certain portion of the cover thickness of these belts is calculated as "carcass" rather than wear cover. With proper design and cover selection, the all-gum belt is suitable for most steel cord belt applications, including some with considerable impact loading.

Figure 3-9 shows the fabric-reinforced steel cord construction, which includes two or more plies of fabric for the most demanding load impact service. The fabric plies carry no belt tension but, rather, are designed with great transverse strength to absorb impact forces and abuse in that direction.

Figure 3-8 All-Gum Steel Cord Conveyor Belt



Figure 3-9 Fabric-Reinforced Steel Cord Conveyor Belt

3. Single Ply Carcass

Single or mono-ply belts are made with a single ply of fabric usually of the types shown in Figure 3-4 and Figure 3-5. The woven fabric is impregnated with an elastomer and then covered with appropriate covers.


B. Covers

Covers protect the carcass of the belt from load abrasion and any other local conditions that contribute to belt deterioration. In a few cases, these conditions may be so moderate that no protection and no belt cover are required. In others, abrasion and cutting may be so severe that top covers as heavy as 12 mm or more are required. In any case, the purpose of cover selection is to provide enough cover to protect the carcass to the practical limit of carcass life.

The pulley side cover is generally lighter in gauge than the conveyor side because of the difference in wear resistance needed. Some belts, however, have the same gauge of cover on each side. Users sometimes turn the belt over when one side has become worn. In general, it is better to avoid inverting the belt because inversion after deep wear on the top side presents an irregular surface to the pulley, which results in poor lateral distribution of tension.


C. Design and Manufacturing Limitations

1. Width versus Thickness

There is no specific ratio of width to thickness that is always good design. However, it can be stated here that the load to be transported governs the width of the belt. If the tension requirement is not the design criterion, then the belt may need investigation on a minimum ply basis to support the load or, in any case, from a maximum ply standpoint to permit troughing.

With steel cord construction, there is more flexibility of design in that the tension member is uniplanar. Proper lateral stability can be independently built into the belt with added rubber thickness in the case of an all-gum belt or with fabric plies on the fabric-reinforced belt.

2. Length of Single Rolls



Conveyor belt roll lengths are limited largely by the ability of the manufacturer, the shipper, or the user to handle the finished roll.

Note that the maximum manufacturing diameter is 4280 mm at the The Goodyear Tire & Rubber Company USA belt manufacturing sites. The probable variance of actual belt thickness from calculated thickness is plus or minus 5 percent. Information on the thicknesses of the various The Goodyear Tire & Rubber Company belts is given in the Product Brochures.

The maximum allowable mass for one roll of belt is 45 metric tons (Marysville, Ohio, plant). As belt roll weights and diameters increase, it becomes more and more important to check for possible limitations that might be encountered either in shipping or in handling the rolls at the job site.

Both mass and diameter maximums sometimes need to be further modified for certain widths and/or types of belt that must be processed through certain equipment that has less than the above maximum capacities.

It is sometimes necessary to determine the length of a roll of belt, especially in the field where it cannot be easily rerolled and measured. Figure 3-12 provides a means of calculating the length.

Figure 3-11 Uncrated Belt Roll Diameters

Figure 3-12 Number of Metres in Roll of Belting



WHERE

L = LENGTH OF BELT IN METRES

A = DISTANCE SHOWN IN METRES

N = NUMBER OF TURNS OF BELT. COUNT TURNS ALONG DISTANCE "C." START COUNT AT INSIDE END OF BELT BUT DO NOT COUNT INSIDE END AS A TURN. BELT SHOWN WOULD HAVE 11 TURNS.

B = ESTIMATED OR MEASURED LENGTH OF BELT IN METRES ON LAST TURN BEYOND POINT AT WHICH COUNT OF TURNS WAS MADE.

NOTE: FORMULA IS INDEPENDENT OF BELT THICKNESS AND SHELL SIZE.


D. Tolerance

1. RMA Standard Width Tolerances

The standard tolerances on width of finished belt as agreed upon by the Rubber Manufacturers Association are presented in Items (1) and (2), at right.

2. Length Tolerance

Length tolerance for belts that are spliced endless at the factory is plus or minus 1/2 percent.

For roll lot belting, no length tolerances are specified but it is accepted by The Goodyear Tire & Rubber Company that, where a roll of specific length is ordered, no minus tolerance is permissible. Only in cases where approximate roll lengths for stock purposes are ordered will any minus tolerance below nominal length be taken.

3. Thickness Tolerance

No thickness tolerances are provided for conveyor belting since thickness variations from belt to belt of a given construction are normally of no consequence. In a few specialized uses of conveyor belting, uniformity of thickness from belt to belt and within a single belt is of importance. In such cases, individual agreement on acceptable tolerance is necessary.

(1) Width Tolerances for Molded Edge Belting

Standard belt width

Tolerance plus or minus

Maximum variation, any

single manufactured

roll of belt 600 mm or less 6.5 mm 6.5 mm

Over 600 mm

1 percent of width

1 percent of width

(2) Cut-Edge Belting

Width tolerance: one-half that shown above for molded edge belting.

Maximum variation in one roll of belt: same as shown above for molded edge belts.

With special attention, closer tolerances can be met. Such requirements should be discussed with the The Goodyear Tire & Rubber Company representative.




E. Molded Belts

1. General

The normal conveyor belt is of rectangular cross section and during vulcanization is formed by what amounts to a temporary mold of rectangular section. This mold is made up of press platens on top and bottom and edge irons of proper thickness at each edge. For any section other than smooth and rectangular, it is not practical to assemble a temporary mold in this way. For such special shapes, it becomes necessary to prepare metal, rubber, or fabric molds that are placed in the press and form the belt to the required shape or surface pattern. The following common molded belts are made by such methods.

2. Raised Edge Type

These belts generally run flat and at slow speed. They carry materials with a high liquid content or they may carry gelatin that is put on the belt in liquid form, allowed to solidify on the belt, and removed in sheets. Metal molds are required for each design.

3. V-Guide Strip

Some conveyors have a V-shaped guide strip vulcanized to the pulley cover. The strip is used for guiding where the belt is transversely out of level, as in certain types of ditching machinery, or is for other reasons not subject to training by normal methods. Metal molds are required for each design.

4. Raised Rib Design (The Goodyear Tire & Rubber Company's Xtra Grip®)

This design is used to increase the permissible angle of incline. Xtra Grip® is especially desirable on materials having high water content where the load might slip on the belt. Gold dredging and wet gravel plants often require such belts. A single standard design is provided from a metal mold (Figure 3-13).

Figure 3-13 Xtra Grip® Conveyor Belt

5. Rough Top Designs

Various designs of rough surface belting, primarily for handling packages on inclines, are molded. While these belts are of rectangular section, the rough surface cannot be molded, practically, direct from the press platens. Molds are placed in the



press for forming the one belt surface. These molds are usually rubber or fabric or a combination of the two and can be made to produce surfaces simulating fabric weaves of various degrees of roughness as well as lightly corrugated or creped surfaces. These designs are pictured in Figure 15-4.

6. The Goodyear Tire & Rubber Company Wedge Grip Design

This diamond-shaped rib-type rough-top design is molded in a continuous curing process (roto-cure) where the design is machined into the center curing drum. This design is also pictured in Figure 15-4.



4-1. BASIC DESIGN CONSIDERATIONS

A. Tension

A conveyor belt is simply a means to an end, a means for transporting material from a start A to an end point B (Figure 4-1).

To perform the work of moving material from A to B, the belt requires power that is supplied by a motor turning a drive pulley. The motor torque translates into a tangential force, called effective tension, at the drive pulley surface. This is the pull or tension required of the belt to move the material from A to B and is the sum of the following:

a. Tension to overcome friction of the belt and conveyor components that contact the belt,

b. Tension to overcome friction of the load, and

c. Tension to raise or lower the load through elevation changes.

Figure 4-1 Conveyor Belt

The relative contribution of each of these to total effective tension varies widely depending on conveyor incline and the load on the belt:

1. An empty belt (level or inclined) has an effective tension consisting only of empty friction (Item a).

2. A loaded level belt will have effective tension consisting of empty plus load frictions (Items a + b).

3. A loaded inclined belt will have effective tension consisting of all three load elements (Items a + b ±c). A slight incline with a light load will be mostly friction (Items a + b) while a steep, heavily loaded belt may be 90 percent or more incline load tension (Item c) with the balance friction. Item c will be plus when material is being elevated and minus when material is being lowered.

In the design of a conveyor belt carcass, it is necessary to determine the maximum tension to which it will be subjected while performing the maximum amount of expected work.

Refer to Section 6 for a complete discussion of belt tension as well as the necessary mathematical formulas and their derivations and usage.


B. Load Support

Tensile strength is not the only consideration necessary in the design of a conveyor belt carcass. Consider Figures 4-2 and 4-3.

Most conveyor belts operate over troughed idlers. The troughing angle of these idlers may vary from 20 to 45 degrees. Obviously, this trough angle affects the belt by creating a line along which the belt is constantly flexed. The greater the trough angle the greater the flexing action. When the belt is fully loaded, the portion of the load (X) directly over the idler angle forces the belt to flex to a shorter radius. The heavier the load, the smaller the radius through which the belt must flex.

Consideration must be given to designing the belt with sufficient transverse rigidity and flex life so that for a given idler angle and load

Figure 4-2 Idler Troughing Angle (20 Deg)

Figure 4-3 Idler Troughing Angle (45 Deg)

Figure 4-4 Belt Design Unaffected by Load Weight

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weight premature belt failure will not occur. This is done by designing the belt with sufficient transverse stiffness to bridge the idler angle within a satisfactory radius.

In Figure 4-4, the belt design is satisfactory in that it does bridge the angle properly under full load. In Figure 4-5 and Figure 4-6, the weight of the load has forced the belt tightly into the angle and premature failure can occur. Other factors that promote this type of failure are belt sag, idler gap, idler design, and idler pitch.

The rubber manufacturers association has issued technical bulletin IP-1-2, date 1995, limiting the idler gap to a maximum of 10 mm. This same idler gap, found in ISO 1537, is entitled "Continuous Mechanical Handling Equipment for Loose Bulk Materials--Troughed Belt Conveyors--Idlers."

Perhaps just as important, the belt in Figure 4-5 will wipe excess grease from the idler bearings. Grease will deteriorate standard rubber compounds and also cause premature belt failure. The properly designed belt in Figure 4-4 bridges this area so that grease will not normally reach the belt.

This design consideration is referred to as load support or the minimum belt construction to bridge the loaded belt over the idler junctions. Actually, in many cases where a belt has been designed for load support, it also will have enough carcass tensile strength to meet the tension requirements.

Therefore, the experienced belt designer often will be able to judge when a load support design will be more than adequate for the tension requirement, thereby eliminating the necessity for tension calculations. The Product Brochures also show load support tables for each constructions. For example, Plylon Plus® data is located here.

Figure 4-5 Belt Design Affected by Load Weight

Figure 4-6 Load Support Failure


C. Troughability

Figures 4-6 and 4-7 illustrate that the belt must be designed to be sufficiently flexible transversely to trough properly.

The empty conveyor belt must make sufficient contact with the center roll of the idler or it will not track properly. In Figure 4-6, the belt is too stiff to contact the center roll and, therefore, will wander from side to side with the possibility of causing considerable damage to the belt edges. In Figure 4-7, sufficient contact is made to steer the belt properly along the idlers.

When designing the belt carcass from a tension standpoint, a check must be made to ensure that troughability design limits are not exceeded. The Product Brochures also show troughability tables for each constructions. For example, Plylon Plus® data is located here.

Figure 4-6 Stiff Belt, Improper Troughing

Figure 4-7 Flexible Belt, Proper Troughing




D. Impact

Another carcass design consideration is impact. The loading arrangement or material being handled may be such that the belt selected for tension or load support considerations would fail prematurely from impact. In such cases, the carcass must be strengthened by using more plies or stronger fabrics.

Impact considerations are detailed in Section 11. The Product Brochures also show impact values for each constructions. For example, Plylon Plus® data is located here.


E. Covers

Once the carcass design is settled upon, attention should be directed to quality and gauge of the covers. The optimum design would mean that the covers would wear out in normal service at the same time as the carcass. This theoretical "total belt" failure after long and useful service will seldom be achieved, but the designer should seek this goal.

Since the carcass of a conveyor belt will fail very rapidly once the covers have worn away, a very small premium for upgrading quality or adding cover gauge may be well repaid in terms of overall belt life. Tables and suggestions are provided as guidelines for the designer, but experienced judgment in this area is the key. The final decision on cover gauge must be tempered by the designer's knowledge of the application, or similar applications.

Complete cover selection is described in Section 8.


F. Other Considerations

The majority of conveyors are relatively simple in design and low in tension so that the foregoing items are all that normally need to be considered in making a belt selection. However, as conveyors become longer, more complex, or higher in tension, it often becomes necessary to investigate one or more of the following:

1. Acceleration and braking problems and tensions (Section 6-11),

2. Coasting time and distance (Section 6-12),

3. Vertical curve tensions and radii (Section 12),

4. Trough to flat transition distances (Section 12),

5. Turnover lengths (Section 12),

6. Two-pulley drive problems (Section 9),

7. Takeup locations and problems (Section 10),

8. Multiple grade conveyor profiles (Section 6), and/or

9. Graduated idler spacings (Section 11).

Attention also is drawn to special belt problems such as package and food belts (Section 15) and grain conveyors (Section 16).

4-2. PROCEDURE FOR SELECTING BELT



A. General

To make a recommendation for a particular installation, the designer has certain data available to him. For a projected installation, the peak hourly tonnage, lump size, material density, and proposed profile are known or estimated. for a replacement belt, not only are these data known, but also the designer is able to take advantage of the knowledge gained from an examination of the belt to be replaced and to determine weaknesses of the old design. Full advantage should be taken of the latter item in particular for determination of the proper carcass and cover for the replacement belt. With this information, by following a logical sequence, the correct belt for the job can be selected. In general, the selection sequence is as follows:

1. Tabulate conveyor data and (when required) select appropriate width and speed.

2. Check for maximum belt tension (Tm ).

a. Experience often indicates no check is necessary, that belt can be selected on a load support basis. Proceed to Step 4.

b. An accurate calculation of Tm is to be used in all cases where tension is a factor in belt selection (see Section 6).

3. List belts that satisfy maximum tension requirements. See Product Brochures for tension ratings.

4. Select those belts from Step 3 that also satisfy load support and troughability requirements (See Product Brochures).

5. Pulleys--Select those belts from Step 4 that also satisfy existing pulley diameters (see Product Brochures).

6. Carcass selection--From the Step-5 list, the final selection depends upon costs, past experience, past belt histories, and impact.

7. Quality selection--Evaluate conditions of heat, oil, abrasion, and cutting; then select the appropriate belt quality (see Section 8).

8. Cover gauge selection--See Table 8-A for a guide to minimum cover gauges.


B. Data Tabulation

A tabulation of all available data is the obvious first step of conveyor belt selection. This may vary widely from the projected conveyor (where the only information may be an approximation of the kind and quantity of material to be conveyed) to the established conveyor (where all or most of the listed data are available).

In the accumulation of data for existing conveyors, the importance of the previous belt history is emphasized. Theoretical and practical design considerations permit precise resolutions of some aspects of the job of designing a replacement belt. Other aspects depend heavily on a complete knowledge of what has happened to the previous belt or belts. This information is absolutely indispensable in making the best possible recommendation for a replacement conveyor belt. This cannot be stressed enough.

Previous belt history is particularly important in determining carcass design for resistance to impact and abuse, as well as establishing cover quality and gauge. In applications that involve heat, oil, and chemicals, previous belt history may be the only immediate source of facts on which to make design judgments.

Again, in any application where a replacement belt is to be recommended, it is incumbent on the designer to obtain



complete PREVIOUS BELT HISTORY.

It is important that a sketch of the conveyor profile be completed. The following must be located by horizontal and vertical dimensions from the tail pulley: drive (each pulley if tandem), beginning and end of vertical curves, vertical curve radii, takeup, change in belt elevation, height, and extent of movement of trippers if applicable (see Figure 4-8). Use of blueprints, profiles, etc., if available, is recommended.

Figure 4-8 Conveyor Profile

DESIGN CALCULATIONS

Max. Running Belt Tension________N kN/m Required________

Belt Required________kW Counterweight Required________N(weight) or _______kg(mass)


C. Width and Speed Selection

For any particular problem of movement of bulk materials by belt conveyor, it is possible to recommend more than one combination of belt width and speed. Lump size permitting, increasing the belt speed permits a decrease in belt width for any specified hourly tonnage. Conversely, an increase in belt width allows a decrease in belt speed. In general, it can be stated that the belt should be kept as narrow as possible (depending on lump size) and run as fast as possible, within accepted limits, to transport the required tonnage. It is also important that the loading chute design be consistent with anticipated belt speed, belt widths, and lump sizes.

The general procedure for selecting belt width and speed is as follows:

1. List material type and size (specified).

2. List material density in kilograms per cubic metre (specified or see Section 5).

3. List peak capacity of belt in tons per hour (specified).

4. See Table 5-C and list maximum recommended speed for material to be conveyed.

The belt speeds in a tabulation of this type will frequently be less than the maximum permissible speed for the material to be conveyed. In Section 6, it is shown that an increase in belt speed at a given loading rate will reduce belt tension so that it may be appropriate to investigate the possible savings in belt cost. However, the following factors often need to be considered in this regard:

1. If the belt construction is already controlled by minimum ply design for load support (Section 7), the reduced tension will offer no savings in belt cost.



2. While the higher speed will reduce belt tension, any savings in belt cost will be partially offset by the fact that the higher speed at a given loading rate also results in a higher power requirement. This will be a cumulative expense over the life of the conveyor as well as a higher initial expense for larger drive components.

3. Increased belt speeds at a given loading rate also will reduce the volumetric load on the belt; i.e., the width of the load will be reduced with the result that cover wear will be less uniform. Also, the decreased time cycle will increase the rate of the cover wear.


D. Typical Conveyor Belt Selection

To illustrate the general process of selecting the proper conveyor belt, consider an incline conveyor with a head drive where tension has already been calculated.

1. Conveyor Data

The primary items of data needed to proceed with the belt selection are:

1. Width: 750 mm.

2. Maximum tension, Tm = 23 400 N = 31.2 kN/m

3. Pulleys: 600-mm head, 500-mm tail and snub

4. Idlers: 35 deg

5. Material: gypsum rock: 1280 kg/m³. Approximately 50 percent fines with lumps to 250 mm

6. Method of joining: fasteners

7. Assume all requirements have been met for normal fastener rating.

2. Carcass Selection

The proper carcass must satisfy all of the following requirements:

1. Maximum tension

2. Minimum plies for load support

3. Maximum plies for empty belt troughing

4. Pulley sizes

5. Impact or other operating requirements.

At this point, a reduction of the available fabric selections is possible simply by general knowledge of the operating conditions (with experience, the designer often will be able to narrow his carcass choices to one or two at this point). In this sample problem, the selection can be narrowed with the following eliminations:

1. All cotton can be eliminated because of the extreme impact abuse of 250-mm lumps. Cotton-nylon



(HDNF) will absorb the distortion and high localized stresses of this type abuse much better than cotton, yet at no increase in cost.

2. Rayon-nylon (HDRN) can be eliminated because the application has a relatively low tension requirement--well below the range normally considered for HDRN.

3. Cover Selection

See Table 8-A for guidance in making cover gauge and quality selection. The following data are required.

Material: gypsum rock-- very abrasive, lump size = 250mm.

With these data, refer to Table 8-A, which is intended to be used only as a guide; its proper use involves careful appraisal of all available service factors and conditions. In other words, the function of the cover is to protect the carcass. Therefore, all knowledge of this and similar applications must be applied to cover selection.

a. Top Cover

Style BII™: 6.5 to 9.5 mm

Stacker®: 5 to 8 mm.

The presence of 250-mm lumps suggests severe cover cutting and gouging from loading impact. With no further specific knowledge of the load point design, the use of Stacker® (RMA Grade I) quality must be considered for its superior cut and gouge resistance. However, if further investigation reveals that the loading chute and skirt boards are well designed, fines are loaded on the belt first, the vertical drop of lumps is held to a minimum, and a loading point is equipped with impact idlers, then Style BII™ (RMA Grade II) covers would provide adequate protection. At this point, knowledge of any previous belt specifications and a performance also would be of great value in making a quality selection.

Without complete knowledge of previous belts and conveyor details with which to make a competent recommendation, the belt designer can only make assumptions and, in most cases, must proceed on the basis of assuming the worst.

b. Pulley Cover

1.5 to 2.5 mm

If previous belt history includes failure due to pulley cover wear, use the heavier gauge. Lacking belt history, use of heavier gauges is also indicated by excessive material spillage, substandard maintenance, and badly pitted idlers.

c. Breaker

The 250-mm lumps dictate that a breaker strip should be included in the top cover for maximum resistance to gouging and cover stripping.

d. Miscellaneous

No other unusual conditions, such as oil or heat, that would affect cover quality considerations are indicated in the data.

4. Final Belt Selection



Combining all developed information yields:

The final selection then would be made based on previously mentioned operating conditions as well as considerations of cost, customer preference, stock availability, and previous belt history.

The above exercise can be accomplished by hand using the Product Brochures, but for a more detailed analysis the The Goodyear Tire & Rubber Company distributor or GTM can provide access to Minuteman System 2000®. This computer program can quickly provide belt and system details to make the belt selection process easier and more accurate. Several screens from this program are provided for reference.

Carcass Top Cover Bottom Cover

Plylon Plus® 625/5 Style BII™ - 6- to 10-mm or Stacker® - 5- to 8-mm 1.5- to 2.5 mm pulley cover

MM1.jpg (27719 bytes)




CAPACITY OF BELT CONVEYORS

5-1. CROSS-SECTIONAL AREA OF LOAD AND TONNAGE CAPACITY--NORMAL MATERIALS

The volumetric capacity of a troughed conveyor belt is determined by the cross-sectional area of the load that can be piled onto the belt without excessive spillage either at the loading point or subsequently due to the small undulations of the belt in passing over idlers. This cross-sectional area is affected by the screen analysis of the material, its moisture content, and the shape of the particles, all of which influence the slope at which the material will stand.

Figure 5-1

Typical Load Cross Section of Normal Bulk Handling Materials on Three Equal-Length Rolls

Since it is usually impractical to evaluate these factors specifically enough to predict their effect on the cross-sectional area of the load, capacity tables are made sufficiently conservative that any ordinary combination of the above conditions can be accepted.

Tonnage capacities for normal bult material on three-roll, equal-length idlers are based on a cross-sectional load area such as that indicated in Figure 5-1. This, of course, does not presume that load shapes are always as depicted here, as they will vary with different materials, dampness, lump size, etc. The load shape is influenced initially by the loading chute and skirtboards. The design of these parts is to some degree controlled by what is expected to happen to the shape of the load after it leaves the skirtboard confinement. However, tonnage capacities derived from this cross-section have been found attainable with most bulk materials and, with favorable combinations of material size and moisture content, loading rates up to 20 percent in excess of these values can be achieved.

To obtain capacities indicated by this method, some normal precautions are required:

1. Free-flowing dry materials or slumping wet mixtures must be considered as special problems. Capacities are determined by methods given later.

2. Lump size limitations tabulated herein must be observed.

3. Skirtboard location at the loading point must be properly designed to give the most advantageous initial load shape.

4. The belt must be trained to enter the loading point centrally.

5. The idler spacing must be suitably related to belt tension to minimize belt sag. This, in turn, will limit load settling and possible spillage.

6. The delivering chute must be pitched (by trial if necessary) to deliver material with a velocity in the direction of belt travel close to that of the belt. This will reduce turbulence and hasten the settling of the load.

7. With lumps near the limit on size, it may be necessary to place lump deflectors on the skirtboards to move inward any surface lumps lying near the edges as the load approaches

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the end of the skirts.

8. The belt capacity thus determined must be considered against peak, not average, requirements.

5-2. CROSS-SECTIONAL AREA OF LOAD AND TONNAGE CAPACITY -- SLUMPING MATERIALS (DRY, FREE FLOWING, OR VERY WET)

Materials that obviously will not stand at the angles assumed for the previous calculations must be handled at a lesser rate to avoid spillage. Such materials are grain, dry sand, cement, mixed concrete, and similar mixtures. Grain capacities are discussed in Section 16. These capacities are based on the assumption that the belt is loaded to within 50 mm of the edge, regardless of width, with a surcharge angle of five degrees.

5-3. CROSS-SECTIONAL AREA OF LOAD AND TONNAGE CAPACITY--WOOD CHIP CONVEYORS

The use of conveyor belts to handle wood chips in the paper industry has found wide acceptance. These conveyors may have more volumetric capacity than the normal bulk conveyor because of the lightness of the material and the fact that the belt can be loaded closer to the edge.

5-4. FLAT AND PICKING CONVEYORS

Flat conveyors are used only to a limited degree for bulk materials due to the loss of capacity when run flat. However, some materials can be removed by plows from a flat belt at any point along its run. It is a fairly common practice to accept reduced capacity in exchange for this simple distribution system in such materials handling operations as foundry sand or wood chip distribution.

For flat belts, it is usual to base capacity on a cross-sectional area one-half that of the 20-degree, equal-roll, troughed belt; hence, capacity tables present in this section can be used for flat belts by taking one-half the tabulated value for bulk materials. This gives a capacity slightly less than the theoretical using a 25-degree surcharge angle and an edge distance (D) = 0.055w + 25)mm., but the error is justified by the obviously greater tendency for edge spillage when the load settles as it moves over the idlers.

Picking conveyors have 20-degree, long center roll idlers to provide wide shallow loads for picking, sorting, inspecting, and feeder applications.

5-5. RATE OF LOADING

The preceding methods of calculating belt capacity give a value that is intended to be a peak rate and not an average for a shift, an hour, or even a few minutes. Even when feed to the belt is controlled by a feeder as it should be, it is not ordinarily safe to assume that the feeder can be set for a rate arrived at by dividing daily tonnage to be handled by hours worked. With the best of scheduling in mining operations, peak tonnage has been found to be 25 percent above the daily average and is often much greater. An estimate of production delays that might result in idle or lightly loaded time for the belt must be made to arrive at a peak rate necessary to compensate for such delays.

5-6. SIZE OF LUMPS

The width of belt required for a material containing large lumps is influenced in two ways by the size of the lumps. First, the cross-sectional area of the load is reduced because the load initially must be kept a greater distance from the edge of the belt. Second, the chute and skirt boards must be wide enough to pass any probable combination of lumps, which in turn sets the minimum belt width, independent of capacity requirements. It happens occasionally that the belt width required to handle lump size is greater than that required for capacity. This condition can only be avoided by crushing or by scalping off large lumps before delivering material to the belt.

Table 5-A shows relationship of belt width to lump size that should be maintained.

Table 5-A Maximum Recommended Lump Size for Various Belt Widths



Belt Width (mm) Lump size (mm)If uniform If mixed with 90% fines

300 50 100400 75 125500 100 150

600-650 125 200750-800 150 250

900 175 3001000-1050 200 350

1200 250 4001350 275 500

1400 and over 300 600

5-7. CAPACITIES FOR SPECIAL CONDITIONS

A. Package Conveyors

Belt conveyors handling large packages, cases, or sacks that feed to the belt singly offer no problems in capacity determination. The belt must be wide enough for the largest package and must run fast enough to take packages away at the peak rate at which they will be fed. There are, of course, limits on speed for package conveyors, sometimes set by the nature of the material and sometimes by the rate at which the packages can be taken away from the discharge end.

Belt conveyors handling packages of miscellaneous size and shape require an analysis of the size distribution of the packages and an estimate of probable combinations of package sizes that might make up a typical cross-section. From such a cross section, belt width and speed can be established. In predicting a typical cross-section, it must be known whether packages can be piled or must be kept single layer for inspection or removal. See Section 15 for more detail on package conveyors.

B. Log Handling Belts

Logs are handled on belt conveyors, both troughed and flat, in pulp wood lengths and in short sawed sticks. Here again it is necessary to analyze the size of the pieces to estimate a typical cross-sectional load. Usually the wood will lie on these belts in a single layer. At some points, the load may occupy practically the full belt width with several small logs side by side, while at others a single log will constitute the load. The capacity given in Table 5-B is based on 60 percent of the belt width and 75 percent of the length being covered with wood averaging 150 mm in diameter. This is a density of loading easily obtained and one that can be exceeded by 35 percent during periods of peak loading.

In calculating tension and power, cords can be converted to tons by using 2000 kg per cord of pulpwood. Then mtph = cords per hour x 2.0. The number of sticks per cord will be between 90 and 125.

Table 5-B LOG HANDLING BELT CAPACITY ( In cords per hour)

Belt Width Belt Speed (mps)0.25 0.5 0.75 1.0 1.25 1.50 1.75

750 12 24 36 48 60 72 96

900 15 29 44 58 72 87 102

1050 18 35 53 70 88 105 123

1200 20 40 60 80 100 120 140

1350 23 45 68 90 113 135 158

1500 25 50 75 100 125 150 175



C. Fully Skirted Belts

With most materials, skirtboards running the full length of a conveyor result in cover abrasion and gouging and for this reason are seldom used for ordinary conveying purposes. They are used on some steeply inclined belts handling lumpy materials that tend to roll back and off the belt. In this case, they are set just off and above belt edge and are not intended to increase belt capacity. Therefore, the normal capacity tables should be used for such conveyors.

Full-length skirts also are used on belt conveyors used as feeders. In this case, the cross-sectional area of load can be increased considerably beyond normal capacity since depth of load can be maintained for the full width between skirtboards, and skirtboards can be set closer than normal to the belt edge. Speeds for feeder belts depend on the abrasiveness of the material but normally do not exceed 0.5 mps for nonabrasive materials and 0.3 mps for abrasive materials.

The depth of the load should not exceed approximately 40 percent of the width between skirtboards. With skirtboards separated 80 percent of the belt width and depth of load at 40 percent of the skirt board separation, the capacity is given by

,

where

T = tonnes per hour (1000 kg tons),

M = weight in kg/m3,

W = belt width in mm, and

S = belt speed in mps.

The general capacity formula for any width and depth of load is

where

A = cross-sectional load area in square meters.

With such heavy loadings as this permits, the idler spacing and belt tension must be investigated to keep belt deflection and idler loadings within acceptable limits and minimize sag and spillage.

To obtain capacities indicated by this method, some normal precautions are required:

1. Free-flowing dry materials or slumping wet mixtures must be considered as special problems. Capacities are determined by methods given later.

2. Lump size limitations tabulated herein must be observed.

3. Skirtboard location at the loading point must be properly designed to give the most advantageous initial load shape.



4. The belt must be trained to enter the loading point centrally.

5. The idler spacing must be suitably related to belt tension to minimize belt sag. This, in turn, will limit load settling and possible spillage.

6. The delivering chute must be pitched (by trial if necessary) to deliver material with a velocity in the direction of belt travel close to that of the belt. This will reduce turbulence and hasten the settling of the load.

7. With lumps near the limit on size, it may be necessary to place lump deflectors on the skirtboards to move inward any surface lumps lying near the edges as the load approaches the end of the skirts.

8. The belt capacity thus determined must be considered against peak, not average, requirements.

5-8. BELT SPEED

Selection of proper belt speed is influenced by capacity required, by the resulting belt tension and power requirement, and by limitations in the nature of the material being handled. Such limitations might be degradation of friable materials, windage losses of light or powdery materials, lump impact on carrying idlers, etc.

As far as capacity is concerned, it is desirable to select a belt speed that will result in a full belt. This produces a better pattern of cover wear. However, it is sometimes necessary to compromise in this respect in favor of belt tension. This is done by increasing speed, which reduces cross-sectional load (with the feed held constant) and thereby reduces tension, permitting a lighter belt. The pin in troughability and saving in initial cost made in this way often offset the loss in cover wear resulting from a less than fully loaded belt.

Speed also has an effect on power requirement, particularly on belts with little or no incline. With tonnage rate held constant, power requirement goes down as speed is decreased. This is because the power to operate the belt and other moving machinery varies directly with speed while the power to move the live load remains constant as long as the rate of loading is fixed. The degree to which speed affects power requirement depends on the ratio of payload to gross load. The higher the percentage of payload, the less effect speed will have on power requirement. In the fairly common case of a level belt with weight of load on the belt equal to the weight of all other moving parts, a 10 percent change in speed would have a 5 percent effect on power. On incline belts, the effect would be less.

Table 5-C indicates the bounds of common practice in conveyor belt speed.

Table 5-C Typical Belt Speeds (Metres per Second)*

Belt Width (mm) Grain or other free flowing material

Run of mine coal and earth (moderately abrasive materials)

Hard ores and stone-primary crushed (very

abrasive materials)300 2.5 1.5 1.5

400 2.5 2.0 1.8

500 3.0 2.0 1.8

600-650 3.0 2.5 2.3

750-800 3.6 3.0 2.8

900 4.0 3.3 3.0

1000-1050 4.0 3.6 3.0

1200 4.6 3.6 3.3

1350-1400 5.0 3.6 3.3

1500-1600 5.0 3.6 3.3



*These speeds are intended as guides to general practice and are not absolute.

Where space limitations and capacity require, the belt speeds in Table 5-C have been exceeded by as much as 25 percent or more in some cases. However, under ordinary circumstances it is better to provide a belt of sufficient width to permit tabulated speed recommendations to be observed. There has been some tendency to limit belt speed when handling friable materials to avoid breakage at the discharge. However, unavoidable vertical drops have been shown to sometimes have a greater effect than belt speed so it may be futile to handicap a conveying problem with extremely low belt speed (1 m/s or less) for this reason alone.

Table 5-D covers special cases of belt conveying not covered by the bulk material speed in Table 5-C.

1800 --- 4.0 3.8

2000 --- 4.0 3.8

Table 5-D Typical Belt Speeds - Miscellaneous Materials and Equipment

Material Belt Speed (m/s)Packages 0.3 - 1.0

Pulp wood (logs) 0.5 (maximum for sorting)

Pulp wood 1.5 - 2.0 (transporting; lower speeds handicap slow discharge)

Car loaders - trimmers - chargers 10.0 - 13.0

Portable conveyors (underground) 1.3 - 2.3

Wheel excavators 4.6 - 5.0

Picking belts 0.3 - 0.5

Belts unloaded by plows 0.5 - 1.0

5-9. WEIGHTS OF MATERIALS

The mass per cubic metre of many materials is subject to considerable variation. The size of material, whether it is wet or dry, and--in the case of minerals--the natural formations account for this variation. Hence, whenever possible, the density for the size and kind of material involved should be accurately determined.

Solid or compact weights, which are available for most materials, cannot be used in the determination of the capacity of belt conveyors and elevators that handle broken or loose materials.

The densities given in Table A-17 in the appendix are based on the normal condition of the material. Table A-17 also includes recommended maximum inclines and preferred cover compounds.



POWER REQUIREMENTS AND BELT TENSION

6-1. CONVEYOR POWER

The power to operate the conveyor is made up of frictional resistance to movement of the various parts, work of elevating or lowering the load, and mechanical losses of the driving system.

6-2. BELT TENSION AND POWER DUE TO FRICTION

A. Friction Factor (C) and Length Factor (Lo)

Frictional loads are due to weight of the belt, weight of the moving parts of the idlers and pulleys, drag of skirtboards and scraper, and drag caused by any minute misalignment of pulleys or idlers.

In addition, the weight of the material on the belt and the internal friction of that material as it shifts and reshapes passing over the idler rolls increase the friction of the system. The calculation of these frictional forces depends upon an assumption of a composite friction factor. The other factors (such as belt weight and material weight) lend themselves to accurate calculation. The accuracy of this assumption, therefore, determines the accuracy of the results obtained in the determination of frictional loads.

Most frictional force components vary directly with the length of the conveyor. However, there are a few components that are independent of belt length and, therefore, can be added in as a constant. The tension and power formulas in this handbook give equally correct values for all center-to-center distances, based on actual data from existing conveyor units used to develop The Goodyear Tire & Rubber Company formula.

From Figure 6-1, it can be seen that the power, by The Goodyear Tire & Rubber Company formula, is proportional to (L + Lo) instead of being proportional to L, where L is taken as the projection on the horizontal of the center-to-center distance (see Section 6-2 C).

The Lo factor is used as a means of including the constant frictional losses, which are independent of belt length and are commonly spoken of as terminal friction. Figure 6-1 shows that The Goodyear Tire & Rubber Company formula gives higher values for short center conveyors and lower values for long center conveyors than does a simple proportion between power and length.

Figure 6-1 Relationship of Conveyor Length (L) to Power Requirements

The composite friction factor used in determining friction force is designated by C. The factor is used in conjunction with a corresponding L factor as indicated in Table 6-A. The value of C depends upon the type of idlers, structure, and maintenance. It also depends on proper relationship between idler spacing and belt tension as a means of limiting internal friction in the load. This relationship is discussed in Section 11, and the C factors tabulated presume its proper use. With belts that require restraint when loaded (that is, that are regenerative), a lower friction factor is shown to place any deviation from the estimated friction factor in the conservative direction (the direction that would indicate larger belt tension and larger power requirement).

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Decline conveyors or decline portions of multiple-grade conveyors that are marginally regenerative should be investigated both as regenerative and as though they required power (that is, by checking with both 0.012 friction and 0.022 or 0.03 as applicable). The most severe results should be used in order to provide a more conservative result.

*The C and Lo factors have proven to be satisfactory for the great majority of conveyor belt tension and power calculations. However, when long, relatively level, heavily-loaded conveyors are encountered where power requirements are large and made up primarily of friction, it is recommended that The Goodyear Tire & Rubber Company be consulted for additional engineering assistance in selecting these factors.

Table 6-A Friction Factor (C) and Length Factor (Lo) for Conveyor Tension Formulas

Class of conveyor Friction factor (C)

Length factor (Lo),

m*

For conveyors with permanent or other well-

aligned structures and with normal maintenance

0.022 60

For temporary, portable, or poorly aligned conveyors.

Also for conveyors in extreme cold weather that are

either subject to frequent stops and starts or are operating for extended

periods at -40 deg F or below

0.03 45

For conveyors requiring restraint of the belt when

loaded0.012 145


B. Q Factor and Belt Weight

1. General

The Q factor represents the weight of the moving parts of the conveyor system and is comprised of the belt weight (B) plus the weights of the moving parts of the idlers. It is expressed as kilograms per metre of center-to-center distance of the conveyor.

2. B and Q Values

Table 6-B gives average values of both B and Q for various widths of conveyors with ply-type belting. These values should not be used with steel cord belting, since both B and Q generally will run much higher. It is recommended that calculating a more accurate Q value in all cases be considered, especially in the following:

1. Level conveyors where power and belt tension are due primarily to friction.



2. Wide belts where estimated values of Q and B can vary widely from actual values.

3. In any case where actual weight of a selected belt varies more than 20 percent from average belt weight above, change Q accordingly and recalculate tension.

4. Conveyors with steel cord belts.

3. Calculation of Q Values

Q can be calculated for any combination of belt and idlers as follows:

where

Q = kilograms per metre;

B = belt mass in kilograms per metre (see Product Brochures);

m1, m2 = mass of rotating parts of carrying and return idlers, respectively, in kilograms (if actual weights are unavailable, see Idler Manufacturers); and

11, 12 = spacing of carrying and return idlers, respectively, in metres.

Table 6-B Average Values of B and Q for Ply-Type Belts (KG/M)

Width (mm)

Light-service material to 800 kg/m³

Medium-service material over 800 to

1600 kg/m³

Heavy-service material over 1600 kg/m³

B Q B Q B Q

300 1 9 2 16 4 25

400 3 12 5 21 6 31

500 6 15 7 27 9 37

600 7 21 9 31 10 43

650 8 22 10 33 11 47

750 9 28 11 42 12 57

800 10 35 12 51 14 69

900 11 39 13 57 16 77

1000 12 45 15 68 19 89

1050 13 49 16 74 21 98

1200 18 60 22 89 27 122

1350 21 74 27 106 33 144

1400 22 82 29 113 36 153

1500 25 92 31 127 40 171

1600 28 105 35 145 45 191



1800 33 131 42 180 54 231


C. Proper Value of Length (L) on Incline Belts

For incline belts, L is the horizontal projection of the center-to-center distance as measured along the contour of the conveyor. For empty conveyors, the amount of belt and equipment to be moved is determined slightly more accurately by the length along the contour of the conveyor.

For a loaded belt, the additional load friction force is determined by the horizontal projection of the center distance. In Figure 6-2,

Figure 6-2 Load Weight Components

T = peak capacity in metric tons per hour,

1000 T = material per hour in kilograms,

S = belt speed in metres per second, and

3600 S = belt travel per hour in metres.

Then:

= material on each metre of belt in kilograms.

and

Also:



= component at right angles to belt.

The latter is the part of the load supported by the idlers. Since T/3.6S is weight of material per metre of belt, the load friction of an incline conveyor is the same as the load friction of a level conveyor having the same length as the horizontal projection of the incline conveyor.

When L is used in calculating the empty belt friction, the resultant empty friction will be lower than actual. However, the small error will be less if L is taken as length along the contour when load friction is calculated. Consequently, this compromise definition of L is justified. Further, on incline belts, friction usually is small compared with incline load tension; hence, error introduced in using a compromise L is small.

Therefore, for convenience, The Goodyear Tire & Rubber Company recommends the horizontal length for both the empty belt calculation and for additional friction power due to the load. The horizontal projection for L does not insert a very substantial error for the average computation. The maximum error is in a belt inclined to the maximum amount (approximately 23 degrees) for the full length, in which case the length along the belt is 8.7 percent longer than the horizontal projection of the length. In the usual case, the difference is considerably less. This 8.7 percent difference enters only into the empty friction estimate and is a much lesser percentage on total effective tension.


D. Components of Belt Friction

Conveyor belt friction is made up of the following:

1. The friction force for the empty conveyor of length L.

2. The additional force to convey the load on the level for a distance L.

These two items are computed separately and are influenced by the type of equipment and the installation (in other words, by composite friction factor C, which is always an estimate).

The power to elevate the load, or the power generated in lowering the load, is independent of the class of equipment or the installation and is subject to exact predetermination.

In the following derivations, the losses in the driving mechanism are not included. The formulas give the amount of power or force that must be provided to the belt at its driving pulleys. Factors used are:

C = friction factor (see Table 6-A);

Q = mass of the moving parts of the equipment in kilograms per metre of center-to-center distance;

L = center-to-center distance in metres or horizontal projection of this distance for incline or decline belts;

Lo = length constant (see Table 6-A);

S = belt speed in metres per second, and

g = gravitational acceleration = 9.8 m/s².


E. Empty Conveyor Friction Force



The empty conveyor friction force can be calculated as follows:

and

In tension calculations, it usually is necessary to determine that portion of the total empty friction occurring on the return side and sometimes that occurring on the conveying side. These forces are calculated as follows in newtons:


F. Load Friction Force

Load friction force can be calculated as follows:

and


G. Fully Skirted Conveyors

There are occasional installations where the load is conveyed between skirtboards over the entire length from the load to the discharge points:

0.2gd2LM = friction force in newtons due to material dragging against the skirtboards

where

d = load depth in meters,

L = conveyor length in meters,



M = material density in kilograms per cubic metre, and

0.2 = constant made up of average material repose angle and coefficient of friction.

To obtain power in kW, multiply this expression by S/1000.

6-3. BELT TENSION AND POWER FOR ELEVATING OR LOWERING LOAD

A. Force or Power Due to Grade

Use T, L, and S as in the previous derivations. Let H be the net change in elevation in metres and A, the angle between belt and horizontal in degrees. In Figure 6-2 then:

In this formula, the tension due to incline load is inversely proportional to belt speed (that is, the greater the speed the less the tension when tons per hour is constant). The power required for elevating the load is independent of the speed and is a function of tons per hour and elevation.

A belt with both incline and decline sections must be investigated under all significant conditions of loading. That is, power or tensions should be calculated for the empty belt; the entirely loaded belt; for inclines only loaded; and for declines only loaded using the appropriate friction factor for each case. Under certain conditions, the belt may generate power and under others may require power. The motor and the belt, therefore, must be selected for the worst condition of loading apt to exist.


B. Additional Force or Power for Trippers

The ordinary conveyor usually discharges over the head pulley. However, it may be necessary to discharge the load at some point before it reaches the head pulley. If such is the case, it is common practice to use a tripper.

Trippers can be either fixed or movable. Both types can be arranged to discharge to either side of the belt or directly back onto the belt. Basically, it is a matter of elevating the load from the normal conveyor level and passing it over a tripper pulley into the discharge chute. The belt then continues on to the head pulley of the conveyor.

Movable trippers are used for stockpiling and can feed directly to the stockpile or bin or to shuttle belts that carry the load away from the main belt on either or both sides. Trippers can be moved by hand, by the power of the conveyor belt, or they can be motor driven. The last two types can be manually controlled or can be arranged to move automatically to distribute the load evenly between the tripper stops.

The tension of the belt at the tripper usually is the maximum in the system due to the elevation of the load. The elevation in a normal tripper is approximately 1.5 to 2 metres, depending on the belt width. Occasionally, trippers for stockpiling have elevated material as much as 15 metres. If more exact tripper dimension information is required, equipment



manufacturers' catalogs will provide it. Since the tension at the tripper usually is high, it is important that pulley diameters be adequate and in accordance with recommended practice.

At the tripper, there are two requirements for power: one to elevate the load, and the other for the friction of the tripper itself and the power to move it when the tripper is moved by the conveyor belt. If the elevation in the tripper has not been included in the net change in elevation for the conveyor, the power requirement should be calculated in the same manner as for the incline load and should be added to the power for lifting originally calculated.

Friction in the tripper itself or for movement of the tripper is small and for belt power calculation can be neglected. A movable tripper, for example, will have the maximum power requirement when the tripper is as close as possible to the head pulley. Since this is not a constant position for the tripper and it occurs only periodically, the tripper friction can be offset by the reduction in power due to that portion of the belt being empty between the tripper and the head pulley.

Movable trippers are commonly operated at speeds varying from 0.1 to 0.3 m/s, although this speed can be changed by appropriate changes in the tripper gearing.

As the tripper is moved opposite the direction of belt travel, the speed of the belt relative to the tripper becomes the sum of the belt speed and the tripper speed. Consequently, the rate of discharge through the tripper chute is increased proportionately; any belts or other conveyors taking the tripper discharge must have capacity for this increased rate.

6-4. EFFECTIVE TENSION, BELT POWER, DRIVE LOSSES

A. Effective Belt Tension and Belt Power

In any belt drive, whether it is transmission, conveyor, or elevator, there exists a difference of tension in the belt on the two sides of the drive pulley. The larger tension is called tight-side tension (T1) and the smaller is called slackside tension (T2). Without slackside tension to prevent slipping, the belt cannot be driven. The difference between the tight side and slackside tension is known as the effective tension (Te), since this tension actually does the work; it is the algebraic sum of the forces just considered:

To this sum would be added the friction between material and skirtboard if the belt is skirted over its entire length.

The last quantity TH/3.6 S is added if the discharge is at a higher elevation than the loading point but is deducted if the discharge is lower than the loading point. In some cases, the discharge is so much lower that Te becomes a minus quantity, which indicates that the force generated by the material seeking a lower level is greater than the total friction losses. In such cases, some means of absorbing the power generated must be provided or the belt speed will become excessive whenever the belt is loaded. Customarily, this power is absorbed electrically as described in Section 9.

The belt power at the drive pulley can be calculated as follows:

The effective tension also can be calculated from watt-meter power readings of actual installations. In this case, motor and drive losses must be deducted from the electrical input to the motor, thus leaving the power absorbed by the belt. Substituting this latter value in the above equations with the proper value of S provides the effective tension. Drive and motor losses must be added to the electrical output of the motor (generator) to obtain the belt power where decline belts



require restraint.


B. Drive Losses Not Transmitted by Belt

In the calculations of belt tensions in this handbook, only the power required at the driving pulley is considered. The friction losses of the terminal pulleys already have been included in the calculations by the Lo factor. It also is possible to estimate terminal pulley friction losses individually.

Methods of connecting the motor and the driving pulleys are numerous. Speed can be reduced through use of belts, chains, gears, enclosed gear reduction units, or some combination of these. In general, the following values can be used in determining the losses in such power trains (Table 6-C).

Losses incurred in couplings, used to control acceleration between motor and reducers, vary roughly from 3 percent to 5 percent. These couplings are electrical or fluid type. Their use and characteristics are detailed in Section 9.

Thus, to determine the actual motor size, it is necessary to add to the calculated belt power the losses caused by the reducers and the coupling. For example, if there are two reductions using open gears, there is a 5 percent loss of input for each reduction. These losses must be applied to the belt power to obtain motor power.

*Add 50 percent to these values in case of exposure to the elements and dusty conditions.

Table 6-C Average Losses, Various Types of Conveyor Drive Reductions

Type reduction Loss (%)

Cut tooth gears, roller chain or belting, each reduction 5

Self-contained spur or helical gears (use manufacturer's efficiency rating) 5

Helical gear reducer coupled with roller chain drive* 5

Herringbone gear reducer coupled with roller chain drive* 5

Worm gear, direct coupled reducer (use manufacturer's efficiency rating) 25

Worm gear, final chain drive 20

6-5. EFFECT OF BELT SPEED

A. On Power

Belt speed does not affect the power required for elevating or lowering the material. The power required to convey the material on the level (load friction) is also independent of speed. However, the power required for the empty conveyor is directly proportional to the speed. Since this power generally is relatively small, the overall effect on the total power is not great. For example, in a typical conveyor, changing the speed from 1.5 to 2.5 m/s increased the power requirement by only 7.5 percent. The belt width and tons per hour remained fixed.

6-5. EFFECT OF BELT SPEED

B. On Tension



When the speed is increased, the empty belt friction tension is not affected. The tension due to load friction and incline load is inversely proportional to the change in belt speed. Using the same conveyor in Section 6-5A where changing the speed from 1.5 to 2.5 m/s increased power 7.5 percent, the effective tension was reduced 35.5 percent. This means that a belt of fewer plies can be used along with smaller pulleys, which is based on the premise that sufficient plies will remain to support the load and that the increase in speed does not adversely affect the material to be carried. For the user, it is therefore necessary to equate the saving in belt and equipment, a one-time saving, against an increase in power cost (a cumulative expense) and against effect of decreased time cycle. In some cases, however, due to maximum tension limitations, an increase in belt speed may permit the use of a single-flight conveyor system where a two-flight system would otherwise be required.

6-6. SLACKSIDE TENSION

Usually, the slackside tension is obtained by a counterweight or by a screw-type takeup. The former is preferable since it maintains a constant tension automatically and can be set at the lowest amount at which the conveyor can be driven. This type maintains a constant tension under all conditions of load, starting, and stretch. Section 10 gives a more complete discussion of both types of takeup.

The amount of the slackside tension necessary is determined by multiplying the effective tension by the drive factor (K). Values of K depend upon the arc of contact between belt and drive pulley (or pulleys), type of takeup, and whether drive pulleys are bare or lagged. Proper values of K are given in Table 9-A and Table 9-B in Section 9.

Calculation of the slackside tension by this method does not in itself necessarily end the problem of slackside determination. Certain belts, such as lowering conveyor with a tail drive, might require more slackside tension than that previously calculated to provide the minimum tension (To) specified in Table 10-C in Section 10. If such is the case, the calculated slackside tension must be increased by an amount necessary to provide the minimum tension (To) at the low tension point of the system. The methods of Section 6-8 (determining maximum tension) show when this is necessary. Conveyors analyzed by the tension diagram method illustrate graphically the location of minimum tension (To).

6-7. TENSION DUE TO WEIGHT OF BELT ON SLOPE

With all incline conveyors, the weight of the belt on the slope causes tension at the top of the slope. This tension can be expressed as BH, where B is the weight of the belt in kilograms per lineal meter and H is the net change in elevation in meters. The formula is derived in the same manner as that for the incline load tension.

In Figure 6-3:

Figure 6-3 Tension Due to Weight of Belt on Slope

and



The extent to which this tension affects the maximum tension is discussed later.

6-8. MAXIMUM BELT TENSION (ARITHMETIC METHOD)

The maximum belt tension is made up of the various components already discussed. For convenience, the formulas for these components are listed as follows:

1. Empty conveyor friction force = gCQ (L + Lo)

a. Return side friction force =gCQL/2

b. Carrying side empty friction = gCQ(L/2 + Lo)

2. Friction force due to load = gC(L + Lo)(100T/3.6S)

3. Total friction force (sum of 1 and 2) = gC(L + Lo)(Q + 100T/3.6S)

4. Incline load tension = gTH/3.6S

5. Effective tension (sum of 3 and 4) = (C(L + Lo)(Q + TH/3.6S) ± TH/3.6S)g

6. Minimum slackside tension = K × effective tension

7. Belt slope tension = gBH

8. Minimum tension = To (Table 10-C in Section 10).

With these formulas, it is possible to determine the tension at any point of any conveyor, but the various tensions must be assembled according to the layout of the conveyor.

Figures 6-4 through 6-12 indicate the composition of the maximum tension for different conveyor layouts. When two formulas are given, both should be computed; the larger value becomes the maximum tension.

The preceding examples provide a method of determining maximum tension for various types of simple conveyors. As the conveyor profiles become more complex with combinations of incline, decline, and horizontal sections, the arithmetic analysis also becomes increasingly complex. It becomes more and more convenient to examine such conveyors either by using the computer as discussed in Section 6-10 or by graphing the tensions as explained in Section 6-10

In making an arithmetic analysis of the more complex conveyor profiles, the following steps normally would be pursued:

1. Calculate Te for all logical loading conditions that might be encountered (empty, full load, declines loaded, and inclines plus horizontal sections loaded). In some extremely long overloaded conveyors, it becomes impractical to assume that all inclines would ever be loaded simultaneously. Such profiles usually are examined for long sections generally inclined, horizontal, or declined that can be assumed entirely loaded even though they may contain small intermediate undulations.



2. From the various Te values, calculate the minimum T2 and counterweight tension required by the drive.

3. Using the counterweight tension from Step 2, calculate tensions along the conveying side for each load condition to the point of lowest tension. If the resulting tension is less than the minimum required (To), then the counterweight must be increased by the necessary amount.

4. Maximum tension may occur at the drive, head, tail, or at some high point along the conveying side. Tension will have to be checked at each probable point for each load condition to obtain maximum tension. After experience, one usually can know in advance the point of maximum tension and the load condition that will produce it.

In calculating tensions along conveyors of this type, it is necessary to add and subtract the various tension components from point to point along both the return and conveying sides (see Figure 6-13).

Figure 6-4 Level Conveyor, Head Drive (Preferred Drive Location)

Figure 6-5 Level Conveyor (Drive on Return Side)

Figure 6-6 Level Conveyor, Tail Drive (Least Desirable Drive Location)



Figure 6-7 Elevating Conveyor, Head Drive (Preferred Drive Location)

Figure 6-8 Elevating Conveyor (Drive on Return Side)



Figure 6-9 Elevating Conveyor, Tail Drive (Least Desirable Drive

Location)

(A) WHEN LOADED BELT GENERATES POWER (PREFERRED DRIVE LOCATION)



(B) WHEN LOADED BELT REQUIRES POWER (LEAST DESIRABLE DRIVE LOCATION)

Figure 6-10 Lowering Conveyor (Tail Drive)

(A) WHEN LOADED BELT GENERATES POWER (LEAST DESIRABLE DRIVE LOCATION)



(B) WHEN LOADED BELT REQUIRES POWER (PREFERRED DRIVE LOCATION)

Figure 6-11 Lowering Conveyor (Head Drive)

Figure 6-12 Lowering Conveyor (Drive on Return Side)



6-9. COMPUTER ANALYSIS OF CONVEYOR BELT TENSIONS

Tensions on simple conveyors can be easily and quickly calculated arithmetically. This approach becomes more and more cumbersome, however, with longer conveyors where there are a large number of incline, decline, and horizontal sections. A computer program is available from The Goodyear Tire & Rubber Company that will take the data for any conveyor profile and print out power requirements and critical tensions for all desired loading conditions.

The computer is programmed to show the power and tensions for the empty conveyor and for the progressively loaded conveyor. In addition, special conditions can be entered into computer. In this example, the computer was programmed to determine the power and tension needed when only the incline was loaded. The highest tension and power requirements occur under this special condition.

Further, vertical curve calculations can be obtained from the computer to assist the design engineer in determining the curve radii needed to ensure proper operation of conveyor system.

Figure 6-13 Adding and Subtracting Various Tension Components



6-10. ACCELERATION AND BRAKING FORCES

A. General

Acceleration or deceleration forces sometimes affect the amount of counterweight required. When such a force passes through the takeup, the counterweight must be heavy enough to hold down the takeup or the force must be limited by lower acceleration or more gradual braking. Otherwise, the counterweight will be picked up, and the slack normally in the takeup will accumulate at some point of lower tension in the system.

Table 6-D shows common conveyor arrangements and indicates those circumstances under which accelerating or braking forces are critical and cannot be ignored.

In some cases, the problem of acceleration is self-correcting. For example, if the counterweight or counterweight travel is insufficient on a head drive incline conveyor having the counterweight following the drive, the belt will slip on the drive



during starting because of lack of slackside tension. This slippage provides the self-correcting feature in that the conveyor accelerates at a lower rate.

A belt may have a sufficient counterweight, but if the counterweight does not have sufficient travel it cannot take care of the added length that is created in the system during acceleration or braking. Therefore, the takeup travel must be in accordance with the recommendations given in Section 10.

The determination of accelerating or decelerating forces is important for conveyors that have concave vertical curves. The radius or length of the curve is dependent upon the maximum tension found in the curve; the greater the tension, the longer the curve. If accelerating tensions are high and not taken into account, the curve will be too short. As a result, the belt will lift off the idlers during starting.

Where the vertical curve already has been established, the alternative to correction of the curve radius is limitation of the rate of acceleration. This limitation would prevent excessive tensions and lifting of the belt in the curve.

*Such takeup problems can be handled by very heavy single counterweight, by use of double counterweight, tail end brakes, or by head and tail driving.

Table 6-D Effect of Acceleration or Braking on Counterweight Takeup

Conveyor geometry Preferred takeup location Acceleration effect Braking effect

(a) Horizontal, head drive Following drive None

Tends to lift counterweight; brakes not

usually large enough to cause trouble

(b) Incline, head drive Following drive or at tail Little or none Little or none

(c) Decline, tail drive At or near head Tends to lift counterweight if decline is slight None

(d) Decline then level portion, tail drive At or near head

Critical; lifts counterweight and feeds slack to foot of

incline*None

(e) Combinations of incline and decline, head drive

Following head or low point in return run Little or none

Critical when stopping with decine loaded; lifts counterweight and slack

runs to foot of the decline*

(f) Same as (e) except tail drive Same as (e)

Critical - lifts counterweight and feeds clack to foot of decline

Little or none


B. Method of Calculation

1. General

With the information in Table 6-D, the point can be determined at which it is necessary or desirable to calculate acceleration or braking forces. These forces can be calculated by the following formula:

where



m = mass to be accelerated or decelerated in kilograms;

a = acceleration or deceleration rate of conveyor in metres per second per second; and

F = force, or tension, in newtons.

2. Weight to be Accelerated or Decelerated

Acceleration and deceleration forces are distributed about the conveyor in direct proportion to the weights of all moving elements in each conveyor section. Weights to be moved include the belt, load, and rotating parts of idlers. Pulley weights have not been included because they have a negligible effect in most cases.

For each significant condition of loading to be investigated on a given conveyor, the total weight to be moved should be calculated as well as the weight on each conveyor section along the top and return runs.

The masses in kilograms per metre of conveyor center distance are calculated as follows:

1. Conveying side only (kilograms per metre)

a. Empty belt: Q1 = B + m1/11

b. Loaded belt: Q2 = Q1 + T/3.6S

2. Return side only (kilograms per metre)

a. Empty belt: Q3 = B + m2/12

3. Total conveying and return sides (kilograms per metre)

a. Empty belt: Q = Q1 + Q3 = 2B + m1/11 + m2/12

b. Loaded belt: Q + T/3.6S

where

T/3.6S = load on belt in kilograms per metre;

m1, m2 = mass in kilograms of rotating parts of top and return idlers, respectively (see appendix)

11, 12 = average spacing in metres of top and return idlers, respectively; and

B = belt mass in kilograms per metre.

With these values, the total weight to be accelerated or decelerated can be determined for any section of the conveyor and under any condition of loading. All that is necessary is to multiply the mass per metre, under the specified loading condition, by the length for the section of the conveyor being considered.

3. External Acceleration Force, Rate, and Time

The starting characteristics of the drive must be programmed to accelerate the greatest power-requiring load condition that can reasonably be expected. For most power-requiring conveyors, this condition will be the fully loaded belt. Where there are combinations of inclines and declines, it is sometimes necessary to select that combination of loaded inclines most



likely to require acceleration and to eliminate some where the likelihood of acceleration is remote.

When an assumed external acceleration force (Fa) is selected to accelerate the maximum power-requiring load condition, it is governed by the following limits:

1. Minimum limit: An acceleration force of 40 percent Te or more should be provided to ensure initial breakaway (total starting force is then 140 percent Te).

2. Maximum limit: The acceleration force is limited so that belt tension during startup does not exceed 150 percent rated belt tension.

An acceleration force of 50 percent is commonly used and falls within these requirements. The total starting force is then 150 percent Te. Since this total starting force is available for other load conditions where Te is less, the accelerating force of those conditions is greater.

With the total motor starting force selected, the accelerating force for each desired load condition can be calculated. If accelerating rate, "a," is needed for other calculations, it can then be obtained from the formula Fa = ma.

If an average accelerating rate, "a," is assumed or provided, then the accelerating force, Fa, is obtained from the formula Fa = ma. A check should be made that the resulting Fa falls between the limits given in the paragraph above. Where a rate, "a," is provided, this rate can be a peak rate rather than an average rate. If the relationship between the peak and average rates is unknown, then the average rate is assumed as approximately half the peak. The average rate is to be used to calculate the accelerated force.

If the time required for acceleration is to be calculated, the following equation applies:

where

t = time in seconds,

S = final belt speed in metres per second, and

a = average acceleration rate in metres per second per second.

4. External Deceleration Force, Rate, and Time

External deceleration forces have no minimum limitation in the sense that the acceleration forces have. In many power-requiring conveyors, there will be no brake and the conveyor simply coasts to a stop with zero external deceleration force. The maximum braking force is limited so that the belt tension does not exceed 150 percent of the rated tension.

The necessary deceleration force, Fd, usually is determined by a prescribed stopping time or deceleration rate. Fd is determined through the following relationships:

and



In a regenerative load condition, the required external braking force must equal the effective tension, Te, plus the decelerating force, Fd. In a power-requiring load condition, the required external braking force must equal decelerating force, Fd, minus effective tension, Te.


C. Effect on Belt Tension - Tension Diagram Method

1. General

An easy and practical manner of determining the effect of accelerating or braking forces on maximum tensions is the use of the tension diagram, whichalready has been explained; this is especially try for conveyors having combinations of incline and decline or some combination of these two with ahorizontal section. In some cases, these forces may be added directly to the running diagram of the conveyor; in others, a freerequired. The following discussion indicates the type of diagram that should be used.

2. Required Type of Tension Diagram

Required types of tension diagrams are:

1. Accelerating Forces

a. Power-requiring load conditions: Add forces to normal tension diagram.

b. Regenerative load conditions: Make a free running diagram to which externally applied accelerating forces are added.

2. Decelerating forces

a. Power-requiring load conditions: Make a free-running diagram to which externally applied decelerating forces areadded

b. Regenerative load conditions: Add forces to a normal tension diagram.

Where the accelerating or decelerating forces can be added directly to a running tension diagram, the effect is to raise the normal baseline by theamount that these forces lie above the normal diagram at the counterweight location on the diagram. If no other point in the system has less than Ttension with this new baseline location, no extra counterweight is required. However, if the relocation of the baseline results in less than Tbetween the baseline and the line of accelerating or braking forces at some low tension point in the diagram, additional counterweight must be added.In addition, a new baseline must be established for the original tension diagram on the basis of the additional counterweight required.

When no extra counterweight is required, the belt is designed on the basis of the maximum tension determined from the normal running tensiondiagram. In other words, momentarily higher tensions, produced by acceleration or braking, are neglected if they do not exceed 150 percent of ratedtension. For the case requiring more counterweight, the belt is designed on the basis of the maximum tension read from the normal tension diagramusing the newly selected baseline imposed by the accelerating or braking forces.

The same comments given in the preceding paragraph apply where the free-running diagram is used.

3. Derivation of Method for Constructing Free-Running Tension Diagrams for Applying Braking or Acceleration Forces

a. General

The basis of this method is a rope passing over a pulley with unequal weights suspended from the ends. If released, the heavier weight will raise thelighter weight with an acceleration influenced by both the sum and the difference of the two weights. A loaded decline conveyor, which will be



accelerated by gravity when its brake is released, behaves similarly. Likewise, a running conveyor being braked to a stop, after power is interrupted,has some similar aspects and is analyzed by similar calculations.

The free-running diagram represents a conveyor without application of any power to any pulley. Consequently, there is no tension change in passingaround the head or tail pulley and the return run tension is equal to conveyor side tension at both head and tail pulleys. Therefore, the height of thediagram is always the same on each side of the center Axis A-A and is also the same at the two outer edges of the diagram. Thus, its purpose is whollyto show the rate of tension change between the head and tail pulley.

b. Gravitational Forces of Acceleration or Deceleration

A regenerative conveyor that has been stopped will accelerate itself without external power due to the effect of gravitational forces of acceleration, FExternal accelerating force applied by the motor is in addition to this natural force.

Similarly, a power-requiring conveyor will come to a stop if power is interrupted due to the effect of the gravitational forces of deceleration, FExternally applied decelerating or braking forces world be in addition to these natural forces.

In each case, the unbalanced force (Fga or Fgd) acting to cause the natural acceleration or deceleration is numerically equal to the effective tension ofthe particular load condition; that is:

This gravitational force is distributed about the various conveyor sections in proportion to the weights to be accelerated or decelerated in the specificsection involved. If the rate, "a," due to these forces is desired, this rate can be calculated from the formula Fga or Fgd = ma. If external acceleration ordeceleration also is applied, the rate due to those forces is added to the rate obtained from the gravitational forces.

c. Assembling the Free-Running Tension Diagram

To assemble the free-running tension diagram, proceed as follows:

1. Plot all gravitational forces (Fga or Fgd) for both top and return runs (see rules algebraic signs of forces, Section 6-10E

2. Above the Fga or Fgd force lines, plot all normal running forces (friction, incline load tension, and belt weight tension). This plottingcompletes the free-running diagram and, if properly assembled, belt tension is the same on each side of each pulley since no externalforces have been applied.

3. Where external acceleration (Fa) or braking forces (Fd) are applied, they are added above the free-running diagram; this results in atension change, which occurs at the pulley where the force is applied to the belt.


D. Effect on Belt Tension - Arithmetic Analysis of Forces

To determine the effects of acceleration forces arithmetically rather than graphically, proceed as follows:

1. Calculate all normal running forces for each loading condition to be considered. Calculate that portion of each force that occurs in each section of the conveyor on both conveying and return sides.



2. Determine the largest counterweight required to satisfy all load conditions considered in Step 1.

3. Calculate the externally required acceleration, Fa, and/or deceleration, Fd, forces that will occur with each load condition. Calculate that portion of each force that occurs in each section of the conveyor on both conveying and return sides by distributing in proportion to the weight to be moved in each section.

4. Where free-running conditions are to be analyzed, calculate the gravitational forces of acceleration, Fga (regenerative belts), or deceleration, Fgd (power-requiring belts). Calculate that portion of each force that occurs in each section of the conveyor on both conveying and return sides by distributing in proportion to the mass to be moved in each section.

5. Start with the counterweight tension from Step 2 calculated for normal running conditions. Add and/or subtract all forces calculated in Steps 1, 3, and 4 in each direction from the counterweight for each load condition being considered (see the rules for algebraic signs of forces in Section 6-10E.).

6. If the analysis of Step 5 reveals no point in the system where tension is less than the minimum required, To, then no additional counterweight is required.

7. On the other hand if there is a point where tension falls below the minimum, then the counterweight must be increased accordingly. This increases tension for all other conditions, including normal running.


E. Rules Governing Algebraic Signs of All Conveyor Belt Forces

Whether forces are to be graphically or arithmetically analyzed, the following rules for their algebraic signs are applicable.

1. Belt slope tension, gBH, and incline load tension, gTH/3.6S:

a. Plus (+) in uphill directions

b. Minus (-) in downhill directions

2. Empty friction [ gCQL/2 and gCQ(L + Lo)] and load friction [ gC(L + Lo)(T/3.6S)] :

a. Plus (+) with the direction of belt travel

b. Minus (-) against the direction of direct travel

3. Externally applied acceleration forces, Fa:


b. Minus (-) against the direction of belt travel

4. Externally applied deceleration forces, Fd:

a. Minus (-) with the direction of belt travel

b. Plus (+) against the direction of belt travel

5. Gravitational forces, Fgd, to decelerate power requirement load conditions:



a. Minus (-) with the direction of belt travel

b. Plus (+) against the direction of belt travel

6. Gravitational forces, Fga to accelerate regenerative load conditions:


b. Minus (-) against the direction of belt travel

Table 6-E Examples of Algebraic Signs of Forces Acting on Conveyor Belts

CONVEYOR SIDE FROM TAIL TO DRIVE

+ gBH

+ gTH/3S

+ EMPTY FRICTION (TOP)

+ LOAD FRICTION

+ Fa (EXTERNAL ACCELERATING FORCE)

− Fd (EXTERNAL BRAKING FORCE)

− Fgd (GRAVITATIONAL DECELERATING FORCE OF LOADED BELT

RETURN SIDE FROM TAIL TO DRIVE

+ GBH

− RETURN FRICTION

− Fa (EXTERNAL ACCELERATING FORCE)

+ Fd (EXTERNAL BRAKE FORCE)

+ Fgd (GRAVITATIONAL DECELERATING FORCE OF LOADED BELT)

CONVEYOR SIDE FROM HEAD TO DRIVE

+ GBH

+ GTH/3.6S

− EMPTY FRICTION (TOP)

− LOAD FRICTION

+ Fd (EXTERNAL BRAKING DEVICE)

− Fa (EXTERNAL ACCELERATING FORCE)

− Fga (GRAVITATIONAL DECELERATING FORCE OF

RETURN SIDE FROM HEAD TO DRIVE

+ GBH

+ RETURN FRICTION

+ Fa (EXTERNAL ACCELERATING FORCE)

− Fd (EXTERNAL BRAKING FORCE)

+ Fga (GRAVITATIONAL ACCELERATING FORCE OF LOADED BELT)



LOADED BELT)

CONVEYOR SIDE FROM HEAD TO DRIVE

+ gBH (L1)

+ (gTH/3.6S)(L1)

− EMPTY FRICTION (L1 and L2)

− LOAD FRICTION (L1 and L2)

+ Fd (EXTERNAL BRAKE FORCES, L2 and L1)

− Fa (EXTERNAL ACCELERATING FORCES, L2 and L1)

+ Fga (GRAVITATIONAL ACCELERATING FORCES, L1 and L2)

− Fgd (GRAVITATIONAL DECELERATING FORCES, L1 and L2)

RETURN SIDE FROM HEAD TO DRIVE

+ gBH(L1)

+ RETURN FRICTION (L1 and L2)

− Fd (EXTERNAL BRAKING FORCE, L1 and L2)

+ Fa (EXTERNAL ACCELERATING FORCE, L2 and L1)

− Fga (GRAVITATIONAL ACCELERATING FORCES, L1 and L2)

+ Fgd (GRAVITATION DECELERATING FORCES, L1 and L2)

L1, RETURN AND TOP SIDE FROM COUNTERWEIGHT TO TAIL DRIVE

+ gBH1 (RETURN)

−BH1 (TOP)

−gTH1/3.6S

+ RETURN FRICTION (L1)

+ EMPTY FRICTION (L1 and L2, TOP)

+ LOAD FRICTION (L1 and L2, TOP)

+ gBH2

+ 100TH2/3S

+ Fa (EXTERNAL ACCELERATING

RETURN SIDE FROM COUNTERWEIGHT TO DRIVE

+ gBH2

− RETURN FRICTION (L2)

− Fa (EXTERNAL ACCELERATING FORCE, L2 RETURN)

+ Fd (EXTERNAL BRAKE FORCE, L2 RETURN)

− Fga (GRAVITATIONAL ACCELERATING FORCES, L2 RETURN)

+ Fgd (GRAVITATIONAL DECELERATING FORCES, L2 RETURN)



FORCES - L1, RETURN; L1 and L2, TOP)

− Fd (EXTERNAL BRAKE FORCES - L1, RETURN; L1 and L2, TOP)

+ Fga (GRAVITATIONAL ACCELERATING FORCES: L1 RETURN; L1 and L2, TOP)

− Fgd (GRAVITATIONAL DECELERATING FORCES, L1 RETURN; L1 and L2 TOP)


F. Effect of Adding External Braking or Accelerating Forces to Gravitational Forces

1. Braking an Inclined or Other Power-Requiring Belt

On interruption of power such a belt would come to a stop by itself, with no externally applied force, because of the action of incline load tension and friction. For a single belt, no brake probably would be required. In a conveyor system consisting of several belts, it may be desirable to provide a brake to prevent dumping material on an earlier stopped belt. Thus, there are two alternatives: (1) no brake but an anti-roll back or (2) a brake to stop the belt more quickly to prevent overrunning and an anti-roll back.

2. Accelerating Decline or Regenerative Belts

When the brake on accelerating decline or regenerative belts is released, the belt is self-accelerated due to the action of the unbalanced force of incline load tension minus friction. The start can be arranged so that the motor is not started until several seconds after the brake has been released. The other alternative is to start the motor the same time the brake is released, thereby bringing the conveyor up to operating speed even more rapidly as the externally applied force would be assisted by the self-accelerating force.

3. Braking Conveyors in Sequence

Where several conveyors operate in sequence, there probably will be some that are multiple grade, some inclined, and some declined. In other words, any particular belt, when loaded or partially loaded, may require or may generate power. This can present a problem in stopping all the belts in the system. For instance if an incline belt follows a decline or regenerative belt, the natural rate (no externally applied force) of deceleration of the incline belt may be greater than the decline belt that is stopped with a brake at an assumed rate of deceleration. This arrangement could cause a pileup of material on the belt that stopped first, in this case the one with incline.

For such conditions, each conveyor that has a natural deceleration should be investigated to determine at what rate each conveyor will stop by itself. The brake size of the overrunning belt must then be selected to stop that belt at an equal or greater rate.

However, if this method of braking control is not used, such an investigation at least warns the user of danger of spillage due to overrun. This investigation also gives him the opportunity of arranging removal of such spillage and providing decking to protect the return run.

This problem is not encountered in starting since the various belts of the system are started in sequence, beginning at the discharge end of the system.



6-11. CONVEYOR COASTING

A. General

When power is interrupted on any power-requiring conveyor, the belt eventually will coast to a stop without application of external braking forces because of the unbalanced forces in the system. Coasting time and distance can be calculated by the methods that follow.


B. Method of Calculation

Coasting times and distances are calculated by application of the following basic formulas:

where

a = coasting deceleration rate in metres per second per second,

t = coasting time in seconds,

d = coasting distance in metres,

Fgd = unbalanced or gravitational decelerating force in newtons,

w = total mass to be decelerated in kilograms, and

S = belt running speed in metres per second.

The unbalanced decelerating force (Fgd) the sum of the following:

1. Effective tension (Te) for the load condition being considered.

2. Frictional losses of the reducer converted to newtons of force. If not known, assume 2.5 percent of the reducer rating. Convert to newtons of force as follows (motor losses generally are small and usually are ignored):

Weights to be decelerated is the sum of the following:

1. Total mass of belt, load, and moving parts for the load condition being considered.



2. WK2 of the drive complex converted to equivalent mass as follows (WK2 is at the high-speed shaft): equivalent mass in kilograms equals


C. Example of Coasting Calculation

With the following assumed data, coasting time and distance can be calculated for a fully loaded belt where

Te = 730,000 N fully loaded

T/3.6S = 450 kg/m

Q = 220 kg/m

L = 650 m

S = 4 m/s

WK2 = 210 kg-m²

Reducer ratio = 37 to 1

Reducer rating = 1000 kW

Pulley diameter = 1600 mm

In calculating the unbalanced decelerating force (Fgd), Te = 730 000 N. The reducer losses can be calculated as:

The total force then is:

Fgd = 730 000 + 4120 = 18 750 = 748 750 N

In calculating the weight to be decelerated, for the load, belt, and rotating parts:

Mass = (Q + T/3.6S)L = (220 + 450) 650 = 435 500 kg

For the equivalent weight of WK2,



Thus, the total weight is

m = 435 000 + 347 609 = 1 782 609 kg

In calculating the deceleration rate (a),

For coasting time (t),

For coasting distance (d),



BELT CARCASS SELECTION

7-1. CARCASS SELECTION, MAXIMUM OPERATING TENSION BASIS

Operating tension calculated for the belt is the first consideration in selection of the proper belt carcass. Other factors such as lateral stability, troughability, etc., are of equal importance, however, and a simultaneous solution providing satisfactory characteristics in these respects is found at the time of selection of the carcass for tension rating.

Initially, the designer knows the characteristics of the material to be transported and the maximum tonnage per hour. The belt speed can be selected using the capacity tables of Section 5. The maximum belt tension can be calculated with these two values and by using the methods of Section 6.

This tension can be expressed in one of two common ways, depending upon the type belt used:

1. For a ply-type belt, it is customary to express the unit tension in kilonewtons per ply per metre (kN/ply/m). A 1200-mm-wide by 6-ply belt having a maximum operating tension of 144 000 N would have a unit tension of 144 kN/1.2 m x 6, or 20 kN/ply/m.

2. Belts made with steel cords are rated on the basis of kilonewtons per metre, as these are single ply insofar as the tension member is concerned. For example, a 1600-mm-wide steel cord belt with 840 000-N tension would have a unit tension of 840 kN/1.6 m or 525 kN/m.

Once the width and the maximum unit tension are determined, it is possible to determine the type belt required. Initially it was stated that, for any given problem of bulk material transportation, there is more than one combination of belt width and speed. In the same manner, once a combination of width and speed has been decided upon and the corresponding maximum operating tension determined, more than one type of belt can be selected. The more moderate the belt tension, the greater is the choice of belt carcasses.

With textiles available today, fabric-type belts can be supplied with tension ratings up to 350 kN/m or more. However, it is often more feasible and economical to consider steel cord belts as tensions increase.

Service factors affecting tension rating recommendations depend on the nature of the belt joint, certain engineering characteristics of the conveyor, and freedom from adverse environment. Where adverse environmental factors are involved, a further service factor is applicable. In this case, the fabric tension rating is to be multiplied by the environmental service factor of 0.75. Adverse environmental factors are defined as:

1. Elevated temperature of the belt reinforcing element due either to high ambient temperature or to the conveying of hot materials. Ambient temperature in excess of 90 degrees C or hot materials in excess of 120 degrees C if fines or 150 degrees C if coarse are likely to result in continuous temperature rise at the reinforcement sufficient to require application of the service factor. This service factor need not be applied when the entire carcass reinforcement consists of fiber or filaments whose strength is relatively unaffected by elevated temperatures, i.e., glass and steel.

2. Slider bed conveyors that, in time, deteriorate the bottom ply to zero strength level as the result of abrasion. An alternative to using the service factor in this case is simply to exclude the bottom ply from the overall evaluation of belt strength.

3. Loss of section not contemplated in original belt design. This could be the result of holes punched for attachment of cleats, etc., or expected service damage to the carcass.

4. Chemical service detrimental to the fiber. Belts for such service may be designed using fiber relatively unaffected by the chemical environment, in which case the service factor does not apply. Where economics or other reasons favor the use of fibers vulnerable to the environment and reliance is placed entirely on exclusion of detrimental material from the carcass, the service factor is applicable.

The lower the tension per mm of width, the greater is the number of possible combinations of plies and fabrics. In such

Página 1 de 57-1 Belt Carcass Selection: Carcass Selection, Maximum Operating Tension Basis

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cases, the belt that will provide the greatest economy of belt and equipment cost should be selected. In the latter item, one choice may permit savings through smaller pulleys and ratios.

Although a belt that was suitable for width and tension may have been selected, it is still necessary to check load support tables to determine if the belt is sufficiently heavy to support the required load. Thus, it may be necessary to use a heavier belt than the tension indicated, especially with heavy materials and large lumps.

With The Goodyear Tire & Rubber Company steel cord belts, the design of the carcass for each width usually provides sufficient load support as well as proper troughing characteristics. For this reason, these belts need not be checked for minimum or maximum ply.

7-2. CARCASS LIMITATIONS, LOAD SUPPORT OR TROUGHABILITY

As described above, certain lightly tensioned belts may have to be designed to satisfactory support the required load. In other words, a belt may provide sufficient strength in tension but, due to size and weight of material, may be too flimsy. Conversely, a belt may become so heavy or thick that it will not be troughable in the width required. The alternative is to use a steel cord belt that will trough in that particular width or use custom manufactured fabric belt with reduced skim coats.

7-3. LIMITATIONS ON WORKING TENSION

A. General

Tension values are only a fraction of the ultimate strength of the materials themselves; to afford some knowledge of factors that require use of a suitable factor of safety, certain conditions that affect working tensions are discussed briefly.


B. Fatigue of Carcass

In any belt installation, the tension member of the belt passes from a higher to a lower or a lower to a higher tension region as it progresses through its cycle. Thus, in a belt system there is a cyclic tension change, high or low in rate, depending upon the belt speed and length.

Laboratory investigation of these repeated stress cycles has shown that when the fabric is stressed to more than approximately 50 percent of its ultimate strength, failure from fatigue occurs very rapidly. The curve of Figure 7-1 shows the result obtained in a test of one particular type of belt carcass.

Figure 7-1 Effect of Tension on Fatigue Life of Belt Carcass


C. Stretch



The more highly a belt is stressed the more it will stretch. The extra stress causes more frequent takeup adjustments for screw-type takeups or more travel for automatic-type takeups. Stretch is based on the belt modulus as determined after 200 cycles of tension according to ISO 283. Stretch can them be divided into permanent and elastic elongation to more accurately predict counterweight movement. See The Goodyear Tire & Rubber Company Product Brochures for additional information.


D. Vulnerability to Accidental Damage

Belts, by the nature of the materials they transport, are subject to accidents. Pieces of tramp iron that pierce the belt, carcass breaks caused by excessively large lumps, cuts, and other operating damages can and do reduce the working cross section of the belt. It is often not convenient to stop and repair a belt the instant it is damaged. Weekend or holidays provide opportunities for scheduled belt maintenance; hence, a belt is expected to run until a convenient stopping time can be arranged.

Here again, it is evident that the carcass must have a suitable factor of safety included in the allowable fabric, cord, or cord stress to provide a reserve for temporary operation under such conditions.


E. Splice Limitations

Conveyor belts can be spliced endless by means of metal fasteners or by vulcanizing. The mechanical fastener has a pull-out strength that is less than the strength of the belt itself. This factor is taken into account by the lower tension ratings provided in the Product Brochures.

Belts of steel cord construction are made endless by vulcanized splicing only. A calculated dispersal of the cut ends and overlapped cords in the splice area makes the strength in the splice theoretically as great as the sum of the strengths of the individual cords in the belt.

Fabric ply belts also can be made endless with a vulcanized splice. While this method does not provide a splice as strong as an undisturbed belt section, it is very strong and is free of local stresses. In any one section of the splice, only one of the plies will be discontinuous; hence, the greater the number of plies, the greater the tensile efficiency of the splice.


F. Pulley Bending Forces

The Product Brochures lists pulley diameters for various belt thicknesses. These diameters were established to obtain the same degree of bending stress, regardless of the belt thickness. No matter what the pulley diameter, the outer plies of a fabric belt must elongate as the belt is bent around the pulley.

The extra stress in the outer ply induced by bending is dependent on the diameter of the pulley, the thickness of the belt, and the elastic constant of the material.

The following example illustrates what magnitude of stresses may be found in the outer plies of the belt while being bent over the pulley:

Assume:

1. Belt is 6-ply.

2. Carcass thickness is 10 mm.

3. Average unit tension is 10 kN/ply/m.



4. Elastic constant, E, is 600 kN/ply/m.

5. Pulley diameters are 900 and 1200 mm.

It is assumed that the neutral plane of a ply belt is through the center of the carcass. Hence, strains in plies above the neutral plane increase and strains below the neutral plane decrease in equal amounts.

Then, unit elongation in outer ply =

(Pulley cover thickness is disregarded.)

For 900-mm pulley, unit elongation in outer ply =

For 1200-mm pulley, unit elongation in outer ply =

Therefore, the bending stresses are as follows:

900-mm pulley: 600 x 0.010 99 = 6.59 kN/ply/m.

1200-mm pulley: 600 x 0.008 26 = 4.96 kN/ply/m.

Using these values, the total stress in the outer ply as the belt is being bent over the pulley is as follows:

900-mm pulley: 10 + 6.59 = 16.59 kN/ply/m.

1200-mm pulley: 10 + 4.96 = 14.96 kN/ply/m.

The Goodyear Tire & Rubber Company has also determined that inner ply compression can also become a problem for the belt when operating at low tensions and/or with small diameter pulleys.

The above calculation and comments explain why the apparent safety factor obtained by comparing tension ratings with ultimate strength is not as ultraconservative as it appears at first glance.

Refer to Face Pressure for a similar calculation for pulley diameter selection on steel cord belts.


G. Lateral Mal-Distribution of Tension

The effect of pulley diameter and belt thickness in increasing unit tension in the outer plies has been shown. A similar effect in the transverse direction is present because of the crown of the pulley over which the belt operates. With a given total tension, the lateral distribution of stress is such that the center of the belt carries more tension than the outer sections



while on and near the pulley. Using the same belt as before and a 1200-mm pulley, 900 mm wide, the following results are obtained:

Assuming standard crown to be 10 mm per metre of pulley face, center diameter = 1200 mm, and edge diameter = 1190 mm, then unit elongation = 10/1190 = 0.0084 mm/mm.. With E = 600, the difference in stress between edges and the center equals 0.0084 x 600 or 5.048 N/ply/m

With an average unit tension of 10 kN/ply/m, the edge tension = 7.48 kN/ply/m, and the center tension = 12.52 kN/ply/m. This increase in unit tension at the centerline of the belt is in addition to the increase in outer ply stress caused by bending.

Similar lateral mal-distribution of forces is found with vertical curves and transition from the last troughing idler to the head or tail pulley. In a concave curve, the center portion of the belt is farther from the center of the circular arc and has a greater tension than the edges. With the convex curve, it is just the reverse and is more severe because the radius of the curve is generally less. This condition is not nearly so severe in a fabric belt as in a steel cord construction because the modulus of steel is so much higher than that of the fabric. These factors are considered in the determination of the radius of the vertical curve as outlined in Section 12.

In the transition from troughed to flat, there are also lateral mal-distributions of tension. To reduce the magnitude of these forces, the belt edges are gradually lowered to the pulley and the center gradually raised (see Section 12). Again, this is more severe in high-modulus belts such as steel cord than in those of low modulus such as nylon.

All of the factors discussed have the effect of increasing unit tension at certain regions of stress concentration beyond the average unit tension that is determined by normal methods of calculation. Some of these factors are subject to fairly accurate determination while others are so dependent on local operating conditions that their effect cannot be precisely allowed for in advance.



BELT QUALITY - COVER GAUGES, HEAT AND OIL PROBLEMS

8-1. GENERAL PURPOSE BELTS

A. Quality Levels

1. General

General purpose belts are belts designed to handle the majority of conveyor applications involving the haulage of normal bulk materials such as sand, gravel, overburden, coal, ores, and rock where special requirements such as heat, oil, and flame resistance are not encountered.

The Rubber Manufacturers Association Industrial Rubber Products Technical Committee is cognizant of the many types of rubber polymers that are now available for use in the manufacture of rubber conveyor belting. Each of these rubbers has characteristics which, when properly utilized, will provide a belt of the lowest cost per unit of material carried under the specified conditions of service.

Many years ago, a range of tensiles and elongations was adopted in the rubber industry for establishing the quality of the grades of covers. At that time, only natural rubber was available so that tensile and elongation were criteria of the quality of the rubber compound.

Today suitable tests have been developed that will evaluate conveyor service characteristics to indicate the general quality of the belt. Efforts are being directed to this end but no single test has yet appeared suitable.

The RMA Technical Committee has, therefore, determined that cover tensile and elongation and cover/ply adhesion can be used to classify belt grades.

For special service involving heat, oil, chemicals, etc., the range and variety of conditions are so great that an investigation of the individual circumstances is required to permit designation of the belt characteristics required. These are discussed below.

2. Grade 1 Conveyor Belting (The Goodyear Tire & Rubber Company Stacker®)

Grade 1 conveyor belting has covers made from natural rubber, synthetic rubber, combinations of natural and synthetic rubber, or combinations of synthetic rubbers to provide the best abrasion-, cut-, and gouge-resisting covers for the specific material and size of material being handled. Grade 1 conveyor belting has a skim coat of rubber compound between the plies; the type of compound used with the particular fabric ensures the highest degree of flexing life. It is recommended for impact and must be used in all cases where cover gauges have not been listed for Grade 2.

3. Grade 2 Conveyor Belting (The Goodyear Tire & Rubber Company Style BII™)

Grade 2 conveyor belting has covers made from natural rubber, synthetic rubber, combinations of natural and synthetic rubber, or combinations of synthetic rubbers to provide the best abrasion resistance. However, Grade 2 belting does not require covers with the degree of cut and gouge resistance of Grade 1.

Grade 2 conveyor belting has a skim coat of rubber compound between the plies; the type of compound used with the particular fabric provides excellent flexing life for normal service conditions where recommended diameter pulleys are used and where the overall operating conditions are less severe than those requiring a Grade 1 belt.

Grade 2 conveyor belting is recommended for impact and for conditions where the Grade 2 and Grade 1 covers recommended are equal in thickness or where the added cover thickness of Grade 2 provides greater economy than Grade 1.


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B. Carcass quality

The quality of rubber needed in the carcass of a conveyor belt is dependent upon the relationship of belt thickness to pulley diameter, tension in the belt, and the frequency of flexing as measured by the time cycle factor (2L/S) where L is the center-to-center length of the conveyor in metres and S is belt speed in metres per second.

In past years, when conveyor belts were made primarily with natural rubber and cotton fabrics, an analysis of each of these items was often necessary because the carcass quality level requirement would sometimes dictate the quality level of the belt. In the interest of manufacturing standardization, where carcass and cover quality needs differ, the overall belt quality was normally governed by the higher of the two.

Today, the cover quality requirements usually dictate the overall quality because, with the compounds and fibers now in use, carcass flex failures are rare.


C. Conveyor Cover

1. General

Ideally, it is desirable to furnish a cover quality and thickness whose service life will match that of the carcass. Since there are so many possible variations of working conditions, such a selection is difficult and often is achieved only after experience with various combinations of carcasses and covers. Certainly, if experience with previous belts is on record, it should not be ignored. Loss of cover by abrasion with carcass relatively intact would indicate that succeeding belts could profitably carry a heavier or a better quality cover. Severe cutting of the cover without serious abrasion loss would indicate a loading problem that would be helped by more cover but probably could be handled more economically by improved loading. Failure due to other causes with cover relatively intact would indicate either less cover or improvement of the conditions producing failure.

To assist the designer in selecting a proper conveyor cover thickness, see Tables 8-A, 8-B, and 8-C. They are based upon the material characteristics, lump size, carcass quality and the time cycle factor (2L/S), or frequency of appearance of any particular section beneath the loading point. While these factors are important, consideration also must be given to other factors that are treated in more detail in Items 2 through 8 below.

Tables 8-A, 8-B, and 8-C are intended strictly as guides in cover gauge and quality selection. More precise tables are not possible because final selections are heavily dependent upon experience, previous belt history, and a general appraisal of service factors encountered by the belt. Final selections may be based on a single dominant factor or on a group of two or more factors. Final gauge selections may sometimes be lighter or heavier than the indicated ranges for the following special reasons:

Factors favoring heavier gauge (higher quality) are low end of time cycle range; high end of lump size range; poor loading (history of cutting, gouging, wear, etc.); ultimate belt life desired (belt is of prime importance; downtime very undesirable); belt in constant use; and costly carcass.

2. Loading Point

The object of any loading point is to introduce material to the belt in such a manner that the least cover wear is experienced. Ideally, material should approach the belt at the same velocity as the belt is traveling and parallel to direction of belt travel. Also, the material should have as little vertical fall as possible. Practically, however, such results are rarely obtained.

Loading on an incline can increase cover wear rate sharply compared to level loading. This is because of load turbulence and slippage as the load settles on the belt and reaches belt speeds. In one particular case where records were kept, a belt loaded on a 14 degree slope had top cover wear roughly three times greater than the belt onto which it emptied with level loading.

3. Shifting of Load



All the wear that occurs on a belt cover is not caused by the speed differential between material being loaded and the belt. This is evidenced by the fact that some long belts show as much cover wear as short belts conveying the same material as a part of the same belt system. The only logical explanation of such a result is that wear also must be caused by movement of the load on the belt as it passes over the idlers.

Idler spacing and belt tension must be kept in proper relation (Section 11), and adequate transverse stability of carcass must be provided ( Section 7) to try to hold this load-shifting wear to as low a value as possible.

Analysis of cover wear on worn belts generally shows cover wear in proportion to load depth; that is, conveying side cover wear follows a pattern similar to the load cross section.

Figure 8-1 is a composite wear pattern measured on the various belts of a conveyor system on completion of the contract. Cover wear varied from a minimum of 0.6 mm to a maximum of 1.1 mm in the individual belts combined for this average wear pattern.

Figure 8-1 Typical Cover Wear Pattern

Table 8-B is intended as a guide to cover gauge selection. A precise table is impossible because final selections are heavily dependent upon experience, previous belt history, and a general appraisal of the factors which follow.

Final selection may be based on a single dominant factor or on a group of two or more factors. Final gauge selection may sometimes be lighter or heavier than the indicated range for special reasons.

Factors favoring heavier gauges are:

1. Low end-of-time cycle range

2. High end-of-lump size range

3. Poor loading (history of cutting, gouging, wear, etc. )

4. Ultimate belt life desired; belt is a prime importance (down-time very undesirable)

5. Belt in constant usage

6. Steep incline

Factors favoring lighter gauges are:

1. High end-of-time cycle range



2. Low end-of-lump size range

3. Good loading (normal cutting, gouging, wear, etc.)

4. Belt of secondary importance; downtime not a serious factor

5. Belt in intermittent usage

6. Basically horizontal

NOTE: There have been a few rough rules for predicting tonnage life of belt covers handling certain materials, particularly bituminous coal, where there is a considerable record of experience available. However, there are so many variables, even within the handling of a single material, that any worthwhile formula for determining tonnage life is next to impossible.

Some typical wear rates measured on RMA Grade 1 as follows in millimetres per million tons:

Coal: 0.025 to 0.050

Overburden: 0.025 to 0.075

Aggregate: 0.038 to 0.064

Ore: 0.075 to 0.125

These, of course, are general figures because of the many variables that affect them. For example, measurements on a coal feeder belt showed cover wear to be as much as 0.17 mm per million tons, which probably is the extreme high-wear rate for coal.

4. Skirtboards and Scrapers

Both of these wear-producing pieces of auxiliary equipment are evils necessary for the operation of most bulk handling belts. However, they are not ordinarily a factor in determining cover gauge and quality. Skirtboards, if not designed to avoid trapping material between the belt and structural parts or if sealing strips are not maintained, can cause severe and rapid wear or gouging of the belt cover. However, correction for this should not be attempted in the belt cover, but in the skirtboards. Guidance on skirtboard design is given in Section 11.

Scrapers often are necessary to clean off material adhering to the belt. heir continual riding on the belt surface is a source of abrasive wear. However, this is ordinarily a minor matter and their use need not influence the selection of cover gauge and quality. Guidance on selection of scrapers and other cleaning devices that will hold cover wear from this source to a minimum is provided in Section 11.

5. Dragging in Spilled Material

Cover wear of this type is definitely controllable with proper maintenance. Periodic clean-up and effective use of belt cleaners can eliminate such wear. Neglect of this factor may cause more wear than the less controllable items of loading point and load shifting wear.

6. Cocked Idlers and Tilted Idlers

These two conditions are conducive to pulley cover wear. Cocking the idlers produces a wiping action completely across the belt as the peripheral travel of the idlers is at an angle to the direction of the belt travel.

Tilted idlers produce a somewhat similar wiping action except that wear occurs only over the troughing rolls, not the center idler roll.



To prevent this type of pulley cover wear, action should be taken to minimize cocked idlers and to limit the degree to which idlers are tilted for belt training purposes.

7. Edge Wear

Edge wear may be of two types: that caused by rubbing against some portion of the conveyor system structure and that which occurs on the conveyor cover near the belt edges as the belt passes over the return idler rolls.

The former is controlled by idler and structure alignment and leveling, by self-aligning idlers on both top and return runs and, as emergency protection, use of lateral limit switches to stop the belt at dangerous extremes of lateral travel.

The latter is caused by build-up of abrasive materials on the rollers near the belt edge. Proper cleaning of the belt and the return idler rolls will alleviate or eliminate this type cover wear.

8. Exposure of the Belt

Rubber exposed to the action of the elements tends to deteriorate. Whether such deterioration will economically justify protection of the conveyor from the elements depends upon the amount of use the belt receives, the belt life expected, the severity of the weather conditions and, as a side issue, value of the protection for the conveyor equipment and the operating personnel. Ordinary conditions of exposure need not influence selection of cover gauge and quality.

Table 8-A takes into account the noncontrollable factors of conveyor cover wear and recommends a proper cover quality and gauge for various materials and circumstances. These presume normal conditions of loading. Also, examination of failed belts may dictate an increase in cover quality or gauge. Such an examination is invaluable in recommending a replacement belt, and full advantage should be taken of it.

Table 8-A Guide to Conveyor Cover Gauge and Quality Selection - Cover Gauge (millimetres)

Moderate abrasive materials

Recommended top cover gauge (mm)Abrasive materials Sharp-abrasive materials

Grain Wood chips

LimeFertilizer

Phosphate rockOverburden

SinterLimestone

Sand and gravelCoal

GraniteQuartz oresTrap rock

Taconite oreCopper ore

All sizes To 75 mm: 3 to 5 To 75 mm: 5 to 61 to 3 >75 mm: 5 to 6 > 75 mm: 6 to 13

Table 8-B Flexsteel® Recommended Cover Gauge (millimetres)

Abrasive materials Sharp-Abrasive materials

Phosphate rock Granite

Overburden Quartz ores

Sinter Trap rock

Limestone Taconite rock

Sand and gravel Copper ore



Coal Top cover 6 to 10 Top cover 10 to 19

Pulley cover 5.5 to 6 Pulley cover 6 to 8

Table 8-C Typical Flexsteel® Cover Recommendations

Service intended Cover compound Cover gauge (mm)

Coal slope MSHA-SBR 10 x 6

Coal overland BMSHA-SBR or Style BII™ 8 x 6

Copper ore slope Stacker® 19 x 6

Copper ore Stacker® 13 x 6

Granite Stacker® 13 x 6

Iron ore slope Stacker® 19 x 6

Iron ore Stacker® 13 x 6

Limestone Stacker® or Style BII™ 13 x 6

Taconite pellets Style BII™ 10 x 6

Waste rock Stacker® 13 x 6


D. Pulley Cover

Pulley cover quality is almost always the same as that of the conveying side. Recommended pulley cover gauges are as follows:

1. Up to 750 mm width: 0.8 to 1.5 mm.

2. 750 mm width and over: 1.5 to 2.5 mm.

3. Fabric-reinforced steel cord belts: 2.5 to 3 mm.

If previous belt history includes failure due to pulley cover wear, use the heavier gauge. Lacking belt history, excessive material spillage on to the return run or substandard maintenance also indicates use of heavier gauges. However, efforts should first be made to prevent spillage and to improve maintenance. Belts with highly stressed carcass such as HDRN or steel cord usually justify the extra protection of a heavy pulley cover.


E. Cover Gauges for Flexsteel® Belts

Table 8-B is a guide for cover gauge selections for the Flexsteel® belt. Gauges indicated in Table 8-B and Table 8-C are substantially greater than those of Table 8-A because they are measured from the cover surface to the cord surface. In the absence of any fabric reinforcement in such belts, the covers are made heavier to provide lateral stability and load support and to give added protection to the vital steel cord layer. This extra thickness protects the cords from cut and gouge penetrations as well as splits through the belt.



8-2. SPECIAL PURPOSE BELTS

A. General

While general purpose belts meet the needs of most conveyors, there are numerous situations where belts may be required to resist the effects of heat, oil, or certain chemicals. Special belts also are required for grain, package, and food handling, but these belts are discussed in other sections of this handbook.


B. Hot Materials and High Temperature Belts

1. General

Elevated temperatures shorten the life of rubber products, and rubber belting is no exception.

On the rubber portions of the belt, elevated temperatures increase the rate of oxidation and continue the original vulcanization so that one of two things can occur:

1. The compound becomes hard and brittle, flakes off, and exposes the carcass, which burns or abrades away.

2. The compound becomes soft and gummy (reverts) and loses its impact and abrasion resistance.

The deterioration of ordinary rubber compounds due to oxidation is generally thought of as doubling with each 10 degrees C rise in temperature.

In today's heat resisting belts, the compounds are made from various synthetic polymers that can be designed to better resist the effects of heat exposure than can natural rubber and still retain respectable abrasion resistance.

The heat resistant belt carcass is, of course, also subject to deterioration when certain temperatures are reached. Figure 8-2 gives the critical temperature areas for some of the carcass materials used in belts.

The carcass often never reaches the temperature of the load or the covers except when running in areas of high ambient temperatures and/or carrying a load of hot fines, which can bake through the cover and into the carcass.

Thus, the covers offer a certain amount of thermal insulation to the carcass, and heat belt covers are usually 3 to 6 mm heavier than normal to increase this protection.

In analyzing a hot materials conveying problem, one of the factors to be considered, other than the abrasive nature of the material itself, is the size. If a material is fine, it can have a more deleterious effect on a belt than a lump material at a given temperature. Fine material makes more complete contact with the belt surface and facilitates condition of heat into the belt cover and carcass. Thus, the belt more nearly reaches the temperature of the material conveyed. Lump material, by its nature, creates voids between it and the belt, allowing some cooling of the belt and the material. Where intermittent lumps are incandescent, as in coke handling, the belt cover may be burned or scorched, but only in isolated spots. Unless repeated frequently, this is a tolerated operating condition.

As thermal conduction is increased with fine materials, the carcass may reach such temperature that its deterioration becomes a primary cause of belt failure.

General classifications of high temperature belt applications are (1) conveying various hot materials in a more or less normal ambient temperature and (2) conveying materials or articles in an elevated ambient temperature.

In (1), the heat enters the belt largely by conduction and radiation from its load. Since the materials of the belt are poor conductors, the heating and deterioration are concentrated in the surface of the belt cover.



Applications similar to (2) are found in the conveying of articles or materials through baking, curing, or drying chambers of varying lengths. Where there is a choice, it is preferable to keep as much of the top run as possible and all of the return run out of the high temperature chamber.

Many such applications are beyond the temperature scope of rubber belts and require various pan type or steel band conveyors.

Figure 8-2 (below) is a general guide to The Goodyear Tire & Rubber Company heat-resistant belt selections. Additional guidance can and should be obtained by maintaining a history of failures of previous belts on any given installation.

2. Heat Belt Selection (Figure 8-2)

Heat belt selection often requires consideration of factors beyond the scope of Figure 8-2. In many cases where there are two or more selections, the belts having higher initial cost may ultimately be more economical because of superior service; this is increasingly true at higher temperatures.

The following factors favor selection of lower initial cost belts:

1. Previous history indicates heat failures are not predominant.

2. Belt used intermittently.

3. Low end of temperature range.

4. Cooler portions of load placed on belt first.

5. Belt is lightly loaded volumetrically and load is laterally dispersed.

6. Load is effectively cooled while on belt.

7. Loaded belt is never stopped. Stopping can destroy a loaded belt in a very short time.

8. Brief time cycle (short dwell time of load on belt).

A significant degree of carcass protection can be attained through the use of extra heavy covers. For conveyors open to the air, each 3 mm of top cover provides approximately 20 degrees F of carcass protection.

The surface temperature of a heat belt is dependent on the type or size of material being conveyed. Baking type loads, or fines, concentrate heat over the belt surface, whereas lumpy material exposes only point contact and allows additional air cooling in void areas. The relationship of material size to compound required is given in Figure 8-2 (below). Cover wear is accelerated by heat, and this alone will justify cover gauges of 25 to 50 percent greater than those subjected only to normal temperatures.

Consult The Goodyear Tire & Rubber Company when service is extremely severe or belt performance is poor.

3. Heat Resistant Belt Comparison and Recommendations (Figure 8-2)

a. Super Thermo-Flo®, Solar-Shield®, Style 6740-A®, and Style BII™ Hot Material

These belts are designed for general industrial hot material conveyor applications covering temperatures from very high (Super Thermo-Flo®) to moderate (Style BII™ hot material). All are designed to resist abrasion, hardening, and cracking when properly applied.

b. Thermo-Chem®



The Thermo-Chem® belt has a heat resistance comparable to Style 6740-A® but is especially compounded for applications such as conveying asphalt, where oil as well as heat is encountered. It is designed to resist the swelling effects of oil as well as the hardening effects of heat.

Caution: Recommended operating temperatures are to be used as a guideline. As maximum operating temperatures are reached or exceeded, the possibility of reduced belt life is increased unless effective measures are taken to cool the belt.

*Fiberglass carcass is designed to resist burn-through up to 538° C.

General Service Legend -- 1 = Excellent 2 = Very Good 3 - Good 4 - Poor

Figure 8-2 The Goodyear Tire & Rubber Company HEAT BELT (Compound Temperature Ratings)

Compound Polymer Carcass Offerings

Service Conditions General ServiceLumpy Baking Enclosed

Heat ABR OilF° C° F° C° F° C°

Solar-Shield® EPDM P/P 400 205 400 205 400 205 1 2 4

Solar-Shield® EPDM Fiberglass* 400 205 400 205 400 205 1 2 4

STF-E EPDM N/N 400 205 400 205 400 205 1 2 4

Thermo-Flo® BUTYL N/N 350 175 350 175 350 175 2 3 4

Style 6740-A® SBR N/N, P/N 350 175 250 125 250 125 2 1 4

Alumina SBR N/N, P/N 350 175 250 125 250 125 2 1 4

Style BII™ Hot SBR N/N, P/N 250 125 200 95 200 95 3 1 4

Thermo-Chem® NITRILE N/N, P/N 350 175 250 125 250 175 2 3 1


C. Oil Bearing Materials

1. General

Several rubber-like compounds have been developed that are more resistant than natural rubber to the action of both petroleum and vegetable oils. In addition to their oil resistant properties they are, as a rule, more heat resistant than the ordinary abrasion-resistant types of natural rubber.

Normally, conveyors that must resist the action of oil do not run in a bath of oil, but carry a product which contains oil in varying amounts. For this reason, the deterioration by oil is less severe than would be experienced in a laboratory swell test in the same oil. The migration of the oils through rubber covers is fairly rapid even at normal temperatures and is accelerated by elevated temperatures. As the oil reaches the carcass, the cover adhesion, as well as the ply adhesion, is reduced, leading eventually to ply separation and belt failure.

Figure 8-3 illustrates the effect of oil on various compounds under accelerated test conditions.



Figure 8-3

Relative Effect of Accelerated Oil Immersion Test on Volume and Hardness of Various Normal Natural and Synthetic Rubber Belt Compounds

The effect of oil on the cover itself is one of swelling, softening, and reduction of tensile. Often, where small amounts of off are involved, the swelling is the primary cause of trouble because it prevents the belt from troughing. Two makeshift remedies for swelling are used. The belt can, in some cases, be inverted periodically to obtain equal swelling on both sides. This fails in effectiveness if the covers are greatly different in thickness. Secondly, the swollen covers can be grooved longitudinally to relieve the lateral compression in the cover caused by swelling. This can be done with a tire regrooving tool. Grooves are spaced closer as severity of swelling increases. Usually, 50 to 75 mm spacing is sufficiently close. Depth of grooves must be sufficient to reach within 1 mm of the carcass.

Improved resistance to the deteriorating effects of oil is obtained with Chemivic®, Neoprene, OMEGA®, MORS®, and Thermo-Chem® belting. The need for such oil-resistant belting is not always recognized at first glance because some materials carry microscopic amounts of oil, waxes, or fats that cannot be seen or felt with the fingers. Investigation of the process through which the material passes may give clues to sources of oil.

The effect of even these minute amounts of oils, waxes, or fats is cumulative on the belt and finally produces the types of deterioration described previously.

Examples of materials carrying oils, waxes, or fats are:

1. Treated coal -- Domestic coal is often treated with relatively small quantities of kerosene-like oils. Coking coal sometimes has small quantities of oil added for processing reasons.

2. Coke -- Partially coked coal, lumps of which are sometimes found in cool regions around the doors of the oven, contain coal tar products that swell rubber compounds.

3. Clay products -- Extruded clay products such as brick and tile carry lubricating oil from extrusion dies.

4. Bakery and other food products contain fats.

5. Materials cleaned by flotation processes sometimes carry oils from flotation medium.

6. Machined parts carry cutting oils.

The synthetic rubber compounds discussed below, are resistant in varying degrees, depending on operating conditions, to various oils and greases.

2. OIL SERVICE COMPOUNDS

a. ORS-Chemivic

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ORS-Chemivic® synthetic rubber is compounded for maximum resistant to mineral, animal, and vegetable oils. It is a blend of The Goodyear Tire & Rubber Company's Pliovic® and Chemigum. Belting with ORS-Chemivic® compound is also recommended for conveying metal turnings where abrasion and cutting are present, and for carrying petroleum coke and crushed soybeans.

b. MORS

MORS® synthetic rubber is made especially for moderately oily conditions such as transport of oily wood chips. MORS® is static conductive to resist static electric charges when the system is properly grounded. Belting with MORS® compound is also ideally suited for handling flaxseed, linseed, cottonseed, kernel corn, and whole soybeans.

c. OMEGA®

Conveyor belting with OMEGA® has moderate resistance to the effect of oils encountered when conveying oil treated materials. OMEGA® qualifies as fire resistant under Mining Safety and Health Administration No. 28-3. OMEGA® has static conductivity of less than one megohm of electrical resistance when the system is properly grounded.

d. Xtra Gard®

This belt is designed especially for use in the grain industry. Xtra Gard® synthetic rubber is ideally suited for conveyor applications carrying flaxseed, linseed, cottonseed, kernel corn, and whole soybeans. Xtra Gard® qualifies as fire resistant under Mining Safety and Health Administration No. 28-3. OMEGA® has a static conductivity of less than one megohm of electrical resistance when the system is properly grounded.

e. WINGPRENE

Conveyor belting with WINGPRENE® synthetic rubber has good resistance to the damaging effects of mineral oils encountered when conveying oil treated materials.

f. Thermo-Chem

Therrno-Chem® is an oil-resistant belt compounded for applications where both heat and oil are a problem. It is resistant to temperatures up to 175° C, abrasion, flexing, oxidation, and the effect of corrosive atmospheres.

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THE DRIVE, PULLEYS, AND MOTORS

9-1. INTRODUCTION

A belt conveyor can be driven at either end or any point along the return run. It is not ordinarily practical to drive along the top run because the presence of the load prevents wrapping the belt around a pulley or pulleys. Without such wrap or other means of providing pressure between belt and drive, it is not possible to apply any substantial pull to the belt.

Driving effort can be applied at more than one pulley and is very commonly applied at two adjacent points as in the tandem or two-motor drive. It also can be applied at two or more remote points such as the two terminals of a conveyor. This technique is commonly considered on long, relatively level overland conveyors, where it can result in reduced belt tension and cost.

Consideration has been given to driving at numerous points along the top run by such means as booster belts or by individually driven idlers. This has not been practiced due to the large percentage of the belt length that must be powered in this way to obtain substantial benefit and to the cost of so applying power.

In general, the recommended drive locations, from the belt standpoint, are as follows:

1. Belts that require power to operate when loaded should have the drive at or near the discharge end.

2. Belts that generate power when loaded and running should have the drive at or near the loading end.

There are, of course, numerous reasons why drives often must be placed at locations other than these. In some extreme cases, the drive may have to be at the opposite end of the conveyor. The result of moving further from the ideal location is:

1. An increasing percentage of total belt length is in high tension, thereby increasing average belt tension.

2. Both elastic and inelastic stretch is greater due to the higher average tension, and takeup travel is increased.

3. Problems with stopping the conveyor may be increased. 9-2. DRIVE TYPES

A. Single Pulley Drive

Single pulley drives, either at the head or the tail, are the most common type. The power that can be applied to the belt at a single pulley is limited only by the tension capacity of the belt. Going to two-pulley or tandem drives, in many but not all cases, saves belt tension. Other advantages of the single-pulley drive often outweigh its cost in belt tension.

The single pulley drive has the advantage of mechanical simplicity. Only the clean pulley side of the belt is contacted for driving purposes. There is less bending of the belt under high tension than with other drives. There are no problems of load distribution between driving pulleys; hence, wear of belt and lagging do not disturb load distribution.

The ability of any drive to transmit pull to the belt depends on are of contact on the driving pulley and on the slack side tension provided for the drive. The relationship between arc of contact and the amount of slackside tension required for a given effective pull is given by the drive factor K, which is defined as the decimal part of the effective pull being transmitted that must be provided as slackside tension. The derivation and determination of values of K are given in Section 2. Table 9-A shows applicable values of K for various conditions of the single pulley drive.

Table 9-A Values of Drive Factor K for Single Drives

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*Column 2 is higher than Column 4 by an amount sufficient to show 20 percent higher T1, with manual takeup. Likewise, Column 3 is higher than Column 5 to

indicate a 20 percent higher T1. This is done to provide a reserve for intermittent and inaccurate tension control that goes with manual takeup and not because

any actual difference in friction between belt and pulley appears when a manual takeup is used. T1, as previously defined, is tight-side tension.

**Column 4 is based on friction coefficient = 0.30.

***Column 5 is based on friction coefficient 0 = 0.35.

Figures 9-1 and 9-2 show the geometry of single drives and, by example, the effect of arc of contact on tight side and slackside tension (T1 and T2).

Figure 9-1 Single Pulley Drive, 180 Deg Arc

Figure 9-2 Single Pulley Drive, 240 Deg Arc

1. Arc of contact

(degrees)

Manual takeup* Automatic takeup

2. Bare pulley 3. Lagged pulley

4. Bare pulley**

5. Lagged pulley***

150 1.20 1.00 0.84 0.67

180 0.97 0.80 0.64 0.50

190 0.91 0.75 0.59 0.46

200 0.85 0.71 0.54 0.42

210 0.80 0.66 0.50 0.38

220 0.75 0.62 0.46 0.35

230 0.72 0.59 0.43 0.33

240 0.68 0.56 0.40 0.30

270 0.58 0.49 0.32 0.24

9-2. DRIVE TYPES

B. Adjacent Two-Pulley Drive



1. General

The principal advantage in the two-pulley drive is that it approximately doubles the total are of contact at the drive. This can substantially reduce slackside tension, and the resulting savings in belt cost might offset the cost of the additional drive. There are cases where no such advantage exists, the most notable being a belt with a long slope and the drive located at the top of the slope. In such cases, it frequently develops that the slackside tension caused by the weight of the belt on the slope is more than adequate for a single drive pulley. Another occasional case arises where the counterweight must be increased beyond the T2 requirements of the drive to reduce belt sag at some low-tension point in the conveyor.

There are two basic types of adjacent two-pulley drives: (1) geared tandem and (2) two-motor. Each of these drives is discussed in more detail after the following discussion of the drive factor (K) and tension distributions in this type of drive.

2. Calculations for Adjacent Two-Pulley Drives

a. Drive Factor (K)

Drive factors (K) are listed for the two-pulley drive and are shown in Table 9-B. These factors, which are minimum values, are valid only when there is an ideal distribution of the total effective belt tension between the primary and secondary drives. For other than this ideal distribution, the K factor, slackside tension (T2), and maximum belt tension all increase. Table 9-C lists K factors for some common two-pulley drive arrangements where ideal effective tension distribution does not exist.

*Tandem drives ordinarily would be provided with lagged pulleys and automatic takeup excepting portable underground equipment where practical considerations dictate a manual takeup. See the comment under single drives concerning coefficient of friction and changes made in K to compensate for the effect of manual takeup on tension.

Table 9-B Values of Drive Factor K for Adjacent Two-Pulley Drives (valid only with ideal Te distribution)

Arc of contact (degrees)

Manual takeup* Automatic takeup

Bare pulley Lagged pulley Bare pulley* Lagged pulley

300 0.51 0.43 0.26 0.19

330 0.46 0.40 0.22 0.16

360 0.42 0.36 0.18 0.13

390 0.39 0.33 0.15 0.11

420 0.36 0.31 0.13 0.09

450 0.33 0.29 0.11 0.07

480 0.31 0.27 0.09 0.06

Table 9-C Dual-Drive K Factors for Some Common Two-Pulley Drive Arrangements (ideal Te does not exist)

Wrap primary drive (deg)

Wrap secondary drive (deg)

Power ratios

1:1 2:1 3:1

180 180 .250 .166 .125

180 210 .192 .128 .124

200 200 .209 .139 .104



b. Tension Relationships, K Factors, and Ideal Effective Tension Distribution

The following symbols are used in this discussion:

T1 = tight-side tension, approaching drive

T2 = tension leaving secondary

K = drive factor, two-pulley drive

Kp = drive factor, primary drive

Ks = drive factor, secondary drive

Te = total effective tension, Tp + Ts

TI = tension between drives

Tp = effective tension, primary drive

Ts = effective tension, secondary drive

Ideal distribution of the total effective tension is achieved, when (1) the slackside tension of each drive is equal to that required by its K value and (2) the high tension side of the secondary drive equals the slackside tension requirements of the primary. When this occurs, the K values of Table 9-B are valid and the following relationships exist:

These two formulas can be solved simultaneously to give the ideal distribution of effective tension:

205 205 .200 .133 .100

210 180 .250 .166 .125

210 210 .192 .128 .096

210 240 .150 .100 .075

220 220 .176 .118 .088

225 225 .169 .113 .085

240 240 .150 .100 .075



The ratio of the ideal distribution of effective tension, Tp/Ts, becomes:

To determine the proper value of K in the adjacent two-pulley drive:

1. Calculate Te total effective tension.

2. Determine primary and secondary power and drive factors Kp and Ks

3. Calculate ideal Te distribution ratio.

If drive motors are to be selected to approximate this ratio or if existing motors are sufficiently large that they can assume this ratio, then use the K values of Table 9-B (above). Note that the effective tension distribution need not necessarily be in the same ratio as motor power when more than calculated kW has been provided.

4. If the motor power distribution is such that Tp/Ts must be less than (1 + Ks)/Kp, calculate drive factor K as follows:

Note: This is derived from the formulas KsTs = KTe and Te = Tp + Ts and provides a higher K value than those of Table 9-B (above) for the common case where T2 must be larger to provide adequate slackside tension for the secondary drive.

5. If the motor power distribution is such that Tp/Ts must be greater than (1 + Ks)/Kp, calculate drive factor K as follows:

Note: This is much less common than the case in Item 4, above, and provides a higher K value than those of Table 9-B for the case where T2 must be larger to provide adequate slack-side tension for the primary drive. This is derived from the formulas



.

c. Examples - Adjacent Two-Pulley Drive Calculations

(1) General

From the foregoing, it is apparent that, by varying the arc of contact and motor power, a nearly infinite variety of K values and T2 requirements becomes possible with two-pulley drives. One thing that is commonly encountered is the desire that all motors be equal for purposes of standardization. This usually forces the power ratio to be either 2 to 1 or 1 to 1, but an infinite variety of arcs of contact still remains, which makes it difficult to set up tables of K values. Sometimes motors are sufficiently oversized that the ideal distribution of effective tensions still can be achieved. The following examples (based on Figure 9-3) illustrate three different situations.

Figure 9-3 Adjacent Two-Pulley Drive

(2) Example 1

Where

S = 1.8 m/s

Te = 45 000 N (81 kW),

Kp = 0.3836 (210 deg)a,

Ks = 0.3836 (2 1 0 deg)a, and

K = 0.083 27a,

Calculate:

1. Tp and Ts, ideal effective tensions at each drive

2. T2, slackside tension

3. TI, tension between drives

4. T1, tight-side tension

5. Select motor power for ideal power split.

aValues of K are used to four places so that calculations using all combinations will check numerically in Example 1. It is not normally necessary to do this.



Because motors are to be selected that will provide the ideal distribution of effective tension for least maximum belt tension, the following procedure can be followed:

The primary belt power equals

and secondary belt power equals

Probable motor selection in standard sizes and to allow for drive losses therefore is:

1. Primary -- 70 kW

2. Secondary -- 20 kW

(3) Example 2

Where

S = 1.8 mps,

Kp = 0.3836,

Ks = 0.3836,

Te = 45 000 N (81 kW)

Motors are: primary--two 35 kW, secondary--35 kW

Calculate:

1. K, two-pulley drive factor

2. Effective tension at each drive

3. T2, slackside tension

4. TI, tension between drives

5. T1, tight-side tension

In this case, it should be seen how nearly the ideal effective tension could be achieved with the existing drives:



The actual motor power split is 2 to 1. However, the 81 kW could be split with a maximum on the primary and a minimum on the secondary to more nearly approximate the ideal ratio of 3.6 to 1 for minimum belt tension. As an illustration, assume 60 kWon the primary and 21 kW on the secondary, or a ratio of 2.86 to 1. This ratio is less than the 3.6 to 1 ideal ratio; therefore, using 2.86 to 1 ratio:

Example 2 requires higher K, T2, and T1 than Example 1 because of the motor power split, although Kp and Ks are the same in both examples. This is necessary to supply required T2 for the secondary drive.

(4) Example 3

Where

S = 1.8 m/s,

Te = 45 000 N (81 kW),

K = 0.08 27 (420 deg total),

T2 = 3747 N, and

Motors are: primary--two 35 kW, secondary--35 kW

Calculate arcs of contact, Kp, and Ks.

In this case it is assumed, for the sake of illustration, that the motors of Example 2 are retained but that the arcs of contact can be varied to provide ideal effective tension distribution and the T1 of Example 1. Because the motor split is 2 to 1, arcs of contact and Kp, Ks values to approximate that ratio need to be found.

If



Kp = 0.663 (150 deg) and

Ks = 0.238 (270 deg),

The ideal Te ratio would be

Then primary belt power equals

(70 kw motor okay).

and secondary belt power equals

(35 kw motor okay).

3. Geared Tandem, Adjacent Two-Pulley Drives

The geared tandem drive is a two-pulley drive using adjacent pulleys driven at a fixed ratio by gears or chains. In the geared tandem, the only means of achieving the desired load distribution is by varying the effective diameters at the two pulleys, by providing a differential between the two drive gears, or by providing a torque-limiting clutch in one of the pulley drives. None of these is very satisfactory and actual distribution of power between the two pulleys is largely unknown in this drive. However, if the slackside tension calculated is provided, and no more, then the secondary drive pulley cannot pull much more than was intended without slipping. Neither can it pull much less than intended without failing to provide slackside tension for the primary drive pulley and causing it to slip. Thus, the secondary drive pulley is self-adjusting at the expense of slip.

Slip tends to wear the lagging down, which is in itself corrective to the load distribution. That is, a single pulley in a tandem drive that is slipping does so because it is trying to carry more than its design share of the load. Wear of its lagging resulting from this slip reduces the effective diameter of the pulley, which reduces its surface speed and, consequently, the share of the load that it attempts to pull.

Many geared tandems have been installed with the two pulleys of equal diameter and have worked, because of the relief provided, as explained above.

Properly, in geared tandems, compensation should be made in pulley diameters for the difference in



distance between bottom surface to pitch line and top surface to pitch line in the belt. A further compensation in diameter should be made for the change in belt speed between primary and secondary pulleys. This speed change is a result of the reduction in tension and consequent contraction in the belt.

The compensation for asymmetry of the belt about its pitch line is made by calculating or measuring the distance from each belt surface to the pitch line. The difference between the two distances then is added to the radius of the pulley running against the side of the belt that has the least distance to the pitch line. Normally, this pulley drives against the bottom side of the belt.

With fabric belts, the pitch line can be taken at the center of the carcass. With steel belts, the pitch line is at the center of the cord.

The compensation for belt speed is made by determining the difference in tension at the points where the belt enters the two drive pulleys (T1-TI). This difference in tension divided by an elastic constant for the belt in question gives unit difference in length of the belt at the two points. A corresponding unit reduction in effective diameter of the secondary pulley then is made.

Using Example 1 shown previously for a tandem drive, assume a 1200-mm-wide, seven-ply belt having a carcass thickness of 10 mm and a cover of 6 mm on top and 1.5 mm on the bottom. The pitch line taken at the center is5 mm + 1.5 mm = 6.5 from the pulley side and 5 mm + 6 mm = 11 mm from the top. The difference is 4.5 mm, which is added to the radius of the primary pulley because that pulley contacts the light cover side of the belt.

Also in Example 1, the tension entering the primary pulley was found to be 48 747 N (T1); tension at the secondary was 13 515 N (TI). The difference in tension is 35 232 N. If the fabric elastic constant is 600 kN/m, then for the belt the elastic constant becomes 600 x 1200 x 7 plies, or 5 040 999 N. Therefore:

(unit length change).

This correction applies only to one load condition and is inaccurate for all others. Hence, it should be made for the most prevalent load condition.

If the nominal secondary pulley diameter were 1200 mm, it would be reduced by 1200 x 0.0070 = 8.4 mm.

With both corrections made on a nominal pulley diameter of 1200 mm, the primary pulley would become 2 (600 mm + 4.5 mm) = 1209-mm diameter and the secondary 1200 mm - 8.4 = 1191.5 mm diameter. These corrections are usually made by ordering lagging of different thicknesses for two identical pulleys.

4. Two Motor, Adjacent Two-Pulley Drives

The foregoing discussion of the geared tandem makes it clear why some measurable method of regulating load distribution between two pulleys of the drive is desirable. The two motor drive provides this means electrically. Each pulley of the drive is driven by a separate motor. One of the motors, usually the secondary, is a wound rotor motor. This permits small adjustments of the speed at which this motor pulls its full load. Thus, the pitchline speed of the secondary pulley can be controlled at the motor rather than by adjusting pitch diameter.

The tension relationship calculations are exactly as shown for the geared tandem. The difference is entirely in the means of producing these calculated relationships. The two-motor drive is so superior in this respect that it has superseded the geared tandem except in portable conveyors underground.



Figures 9-4 through 9-8 show the distribution of load in a two-motor or tandem drive as affected by the symmetry or asymmetry of the belt and differences in pulley diameters before correction by adjustment of motor speed or pitch diameters.

In all these cases, it can be seen that slip on the primary or secondary pulley finally limits mal-distribution of load, in a tandem or two-motor drive, caused by failure to correct effective pulley diameter or by improper correction. In the geared tandem, this slip persists and wears the pulley lagging. In the two-motor drive, the forces tending to cause slip are transferred to the motor and produce electrical slip. At this point, its effect on load distribution is mitigated by varying the secondary resistance of the wound rotor motor.

Figure 9-4 Symmetrical Belt (Pitch Line Equidistant from Top and Bottom Surfaces), Equal Pulley Diameters

Figure 9-5 Assymmetrical Belt (Pitch Line Raised on Primary), Equal Pulley Diameters

Figure 9-6 Assymmetrical Belt (Pitch Line Lowered on Primary), Equal Pulley Diameters

Figure 9-7 Symmetrical Belt (Primary Pulley Diameter Too Large)



Figure 9-8 Symmetrical Belt (Secondary Pulley Diameter Too Large)

*Sufficient slackside tension was provided in Figures 9-4 through 9-8 permit the secondary to pull half the full load without slip, which is the reason these figures approach 50-50.

5. Two-Motor Drive at Remote Points (Head-Tail Drive)

The preceding discussion of two motor drives is based on the two motors being immediately adjacent to each other. On some conveyors, such as very long, relatively level hauls, there is a further saving in belt tension available by separating the two motors, for example, one at the head and one at the tail Figure 9-9).

Figure 9-9 Head-Tail Drive

If drives are placed at head and tail as illustrated in Figure 9-9, ideal distribution of total effective tension for minimum belt tension is calculated by solving the following simultaneous equations:

(9-1)

, (9-2)

where



Kp = drive factor at head (primary) drive;

Ks = drive factor at tail (secondary) drive;

Te = total effective tension;

Tp = effective tension, head (primary) drive;

Ts = effective tension, tail (secondary);

friction tension on return run between drives;a

gBH = belt weight tension between drives;a

change in belt tension along return run between the drives.a

The following example illustrates a typical reduction in belt tension by using a head-tail drive versus an adjacent two-pulley drive:

Assume

Te = 350 000 N,

gBH = 2500 N,

= 70 000 N,

Kp = 0.38 (head end), and

Ks = 0.38 (tail end).

Substituting in Equations 9-1 and 9-2 and solving for ideal Tp and Ts,

350 000 = Tp + Ts

0.38 Tp = 1.38 Ts + 2500 - 70 000

Ts = 113 920 N

Tp, = 236 080 N

Tm = (1 + Kp)Tp = 1.38 (236 080) = 325 790 N,

where



Tm = maximum belt tension.

If the two drives had been adjacent at the head end with an ideal K value of 0.09, the maximum tension would have been 1.09 (350 000), or 381 500 N. By using a head-tail drive, a tension reduction of 381 500 - 327 590 = 55 710, or approximately 15 percent, was realized.

The advantage of the drive at head and tail decreases as incline becomes greater and is nil when gBH becomes equal to gCQL/2; i.e., when there is no change in tension along the return run. This brings out the point that the advantage of the head-tail drive is due to the tension change between the two driving pulleys resulting from idler friction. When there is no substantial tension change, there is no advantage in head and tail driving.

On decline conveyors with small grades and the drive normally at the head, there is a percentage saving by driving at head and tail but, quantitatively, the saving is likely to be small because such conveyors do not require much power.

On declines with heavy grades, where the drive is normally at the tail, there is seldom any saving in head and tail driving over a two-motor drive at the tail.

Just as with adjacent two-pulley drives, if motors are not selected that will give the calculated distribution of effective tension, then counterweight tension and maximum tension will be increased due to the higher T2 requirements of one of the two drives.

It also should be noted that further tension reduction would be possible by increasing the arcs of contact. It is not uncommon, for example, to install an adjacent two-pulley drive at the head and a single drive at the tail. On very long belts, the additional belt savings often will pay for the added drive.

a In these formulas, L is the length along the return run between drives and H is the elevation change between drives. This is for the case where one or both drives may not be at or relatively near the terminal pulleys.

9-2. DRIVE TYPES

C. Miscellaneous Drives

1. General

In the preceding drives, pressure is produced against driving pulleys by the tension in the belt. Tension capacity in the belt is one of the factors in its cost; hence, efforts to save on tension are logical. Tandem and two-motor drives have been effective ways of saving tension, but they still use belt tension to produce pressure. They are good because they use only very little tension beyond the effective pull, which is unavoidable.

To obtain pressure against the pulley without expending tension in the belt, other drive devices have been invented and, to some degree, put into use.

2. Hugger Drive

One such device is the hugger drive (Figure 9-10), which employs a separate tension belt to produce a pressure on the main belt and the drive pulley. The capacity of this drive is limited only by the tension capacity of the hugger belt; it could operate with zero slackside tension. However, some slackside tension is always necessary for other than driving reasons so the net saving in tension between hugger and two-pulley drives is small. Such a drive has a complication of pulleys and a maintenance problem on the tension belt.

Figure 9-10 Hugger Drive



3. Mangle Roll Drive

The mangle roll drive (Figure 9-11) produces pressure between belt and pulley by a squeeze roll. The capacity of the drive is limited by squeeze roll pressure and arc of contact preceding the squeeze roll. It could also operate with zero slack-side tension. This drive has not been used on bulk material conveyors but has some application to package conveyors where there is no problem of dirt or material getting into the pressure section under the mangle roll.

Figure 9-11 Mangle Roll Drive

9-3. PULLEYS

A. General

Pulleys are one of the most important components of a conveyor system as they are required every time the belt direction changes. They are also a critical component of every drive. A pulley is a system of sub-components consisting of a pulley, shaft, locking device and bearings. It is a generally accepted engineering practice to consider all of these parts together as a composite structure.

Figure 9-12 Photo Courtesy Dodge Divison Rockwell Automation



9-3. PULLEYS

B. Construction

A pulley is made up of many components.

The shaft diameter required is a function of strength and shaft deflection. Shafting materials can be AISI C1018 or C1045 steel.

The pulley itself is constructed from a rim, end disk and hub. It may or may not have one of multiple center disks to support the rim. Welded connections are made between the rim and end disk and between the end disk and hub.

A locking device or bushing is used to secure the pulley to the shaft.

Figure 9-13 Photo Courtesy Dodge Divison Rockwell Automation

9-3. PULLEYS

C. Types

By far, the most commonly used pulley is a standard drum (Figure 9-14). They are built in a wide variety of width and diameters. They are used as drive, takeup, tail, snub and bend pulleys. They are commonly lagged with rubber to increase friction on driven pulleys and increase wear resistance on non-driven pulleys. The rubber lagging also helps shed material buildup to help prevent damaging the conveyor belt.

Figure 9-14 - Standard Drum Photos Courtesy Dodge Divison Rockwell Automation



Standard welded pulleys are defined by the American National Standards Institute (ANSI) standard number B105.1. This standrad establishes load ratings, nominal dimensions and premissable crowns.

Standard pulleys should not be used with steel cord or other high modulus conveyor belts. These high strength/low stretch belts create higher stress concentrations and require more exacting tolerances beyond standard pulley specifications. Starting, braking and and other dynamic loads ar more directly transmitted to pulleys therefore concentricity, manufacturing tolerances and installation precision are more important. These belts require engineered class pulleys. An engineered pulley is one which has been specifically designed to meet load conditions of a particular application.

Wing pulleys (Figure 9-15) are typically used as tail pulleys or in other dirty areas of the conveyor where trapped material is likely or probable. If material gets into this pulley, it will fall between the wings and fall out the sides of the pulley.

Figure 9-15 - Wing

Spiral wing pulleys (Figure 9-16) are a variation of a wing pulley with a spiral wrap. The spiral wrap also allows trapped material to fall through. However the spiral prevent the vibration or "knocking" inherent with a standard wing design.

Figure 9-16 - Spiral Wing



9-3. PULLEYS

D. Pulley Diameters

Pulley diameter is related primarily to belt thickness, belt tension, bending stress induced in the outer plies, and face pressure between belt and pulley.

In the recommended pulley diameters, all these factors have been given weight. For steel cord belts, face pressure alone is the determinant on drive pulley size at the higher tension ratings.

9-3. PULLEYS

E. Pulley Crown

The standard pulley crown is a taper of 10 mm per meter of face width. This results in an increase in pulley diameter, at the center, of 10 mm above the edge diameter for each metre of face width; i.e., a 1-m-wide pulley will be 10 mm larger in diameter at the center than at the edges.

The practice of increasing crown with pulley width is wrong in certain respects. It is more proper to increase crown with diameter and hold it constant for all widths. Furthermore, a crown is most effective when it has a long unsupported span of belt approaching the pulley. The lateral position of this span then can be influenced by the location of the belt on the crown without resistance being offered by the supporting idlers. In conveyors, this condition does not exist on the top run and, consequently, crown on the head pulley is of little value in training the belt. It is a distinct detriment as far as lateral distribution of tension in the belt is concerned.

Head pulleys, therefore, should be uncrowned in normal circumstances. For tail pulleys and takeup pulleys, which normally have a fairly long approaching span without support, crown is moderately beneficial.

Table 9-D shows where the crown may be used, but flat pulleys are preferred in all cases from the belt standpoint since the limited benefits of crown can rarely justify the resulting stress mal-distributions.

Table 9-D Permissible Pulley Crown

Type Belt

Level or inclined conveyors - drive at head or on return run Decline conveyors - drive at tail or on return run

Head or other high-tension

pulleysTail pulley Takeup and

snubsTail or other high-tension

pulleysHead pulley Takeup and

snubs

Fabric - tension at or above 13

None 10 mm/m* 10 mm/m* None 10 mm/m* 10 mm/m*



*Flat pulleys are preferred in all cases to prevent stress mal-distribution.

kN/ply/m

Fabric - tension under 13 kN/ply/m

10 mm/m* 10 mm/m* 10 mm/m* 10 mm/m* 10 mm/m* 10 mm/m*

Steel cord Belts None None None None None None

9-3. PULLEYS

F. Face Width of Pulleys

It is common practice to make the pulley face 50 mm greater than belt width for widths through 1000 mm and about 75 mm greater for widths more than 1000 mm. However, The Goodyear Tire & Rubber Company recommends increasing these standard values to 75 and 100 mm, respectively, as a protective measure for the belt. For long, highly stressed, costly belts and all steel cord belts, the insurance against edge damage to the belt provided by extra face width on the pulleys is well worth its cost. For such cases, the next larger standard face width should be used.

9-4. PULLEY LAGGINGS

A. General

Lagging on drive pulleys is normally used to improve the coefficient of friction between belt and pulley. This is reflected in the values of drive factor K previously tabulated. Laggings also are used on idle pulleys because they are to some degree self-cleaning and because they prevent wear of the pulley surface. Incidental but valuable uses are tapered laggings to reduce or erase crown and diameter correcting laggings to adjust two-pulley drives to the pitch diameters required for proper load distribution.

Laggings can be vulcanized to the pulley surface, in which case there is no fabric reinforcement; they can be cemented to the pulley; or they can be bolted to the pulley using flat head elevator bolts. In this case, a fabric carcass, usually not more than four plies of any belt duck, is used to hold the bolts and resist the tension under which the lagging is applied.

Bolted-on laggings are made without rubber cover on the inner side and with 3 to 6.5 mm cover on the upper side. Where no grooving of the surface is intended, 3 mm cover is a proper minimum. If grooving is to be provided, 6.5 mm cover is a necessary minimum.

Bolted lagging is usually applied to the pulley in two or more circumferential strips. It can be handled more easily this way and applied with higher unit tension. Also, the joints can be staggered around the pulley circumference. For idle pulleys, a lagging is sometimes applied by spiraling a narrow strip onto the pulley, bolting only at the ends of the strip, and using an air-curing cement on the balance of the lagged area.

For vulcanized-on lagging, the pulleys must be sent to the rubber manufacturer who applies, usually, a 10- to 15-mm thickness of rubber, using proper bonding practices to vulcanize it in place. Such laggings are not usually ground after vulcanizing as they can be built up with sufficient accuracy for the purpose without that added cost.

Vulcanized-on lagging is available in various degrees of hardness depending on the purpose. For driving purposes, a compound with 55 to 60 Shore hardness is recommended. For idle pulleys where belt tension is low and lagging is used for cleaning only, the hardness should be low, i.e., 40 to 45 Shore hardness. Grooving the rubber to permit greater deflection of the surface improves the ability of the pulley to clean itself.

Cemented-on laggings using self-curing cements have achieved results comparable to the vulcanized laggings with the added advantage of field application. Great care must be taken, of course, in the use of proper cements and procedures to ensure maximum adhesion between lagging and pulley surface.




B. Grooved Lagging

Grooved pulley lagging was first developed by The Goodyear Tire & Rubber Company for the long decline belts at Shasta Dam. Here, the heavy rains falling on the unprotected return run resulted in a very wet pulley surface of the belt as it approached the tail drive. Because the drives were at the tail, no idlers contacted the inner side of the belt to roll off the water.

Tests on a dynamometer showed that, with a properly grooved lagging, water could be squeezed out and a driving capacity equal to that of dry belt on smooth lagging could be obtained. The actual comparison is shown in Figure 9-12, which graphically represents the resistance to slipping of both dry and wet drives with both smooth and grooved laggings. These test results were confirmed by field performance.

Figure 9-12 Effect of Grooved Pulley Lagging on Slip

Where there is no problem of wet drives, grooved lagging is sometimes justified by its increased deflection under belt pressure, which helps to break dirt accumulations.

Grooving of lagging can be done either on vulcanized or bolted-on lagging. The pattern of the grooving is usually a chevron with the apex pointing in the direction of belt travel to squeeze out the lubricating fluid (water and slimes). The higher the viscosity of this fluid, the more extensive grooving should be. In the extreme cases, it may be necessary to add water to get the mixture to a fluidity that can be squeezed out. Figure 9-13 illustrates one method and shape of grooving.

Grooved lagging adds nothing to the coefficient of friction between dry belt and pulley as Figure 9-12 shows; hence, its use is not a valid reason for further decrease in the drive factor K shown for lagged pulleys.

Figure 9-13 Grooved Pulley Lagging




C. Ceramic

Ceramic lagging has been used for many years where very high friction is required or where lagging wear is extreme. Initially, ceramic was molded into the exact shape the outside diameter of the pulley rim and bolted into place. Although this lagging will not wear and does produce very high friction values, the cost is very high and the lack of give in the ceramic can lead to problems with material buildup and even lagging damage.

Recent lagging developments include combining of best properties of rubber and ceramic. Ceramic chips are imbedded into a rubber matrix. The belt contacts the ceramic chips thereby producing high friction and high wear resistance, but the chips are able to float in the rubber matrix therefore eliminated the original ceramic problems.

9-5. THE POWER UNIT

A. General

Belt conveyors are almost universally driven by electric motors; therefore, no discussion of other drives is provided. AC squirrel cage induction, AC wound rotor induction and DC motors are the three types most widely used for conveyor belt drives. Those motors are generally available in the following standard power sizes:

1. 1, 1-1/2, 2, 3, 5, and 7-1/2 kw

2. 10 to 25 kw in 5-kw increments








The following information will be subdivided into three main sections based on the three principal motor types. The information by no means affords sufficient data for the design of the drive complex. More complete information on drive design should be obtained from the equipment manufacturers. The chief concern for the belt designer is that the starting time be of sufficient duration as to limit the belt tension to 150 percent of normal tension rating during the starting period.

9-5. STARTING TYPES

B. AC Squirrel Cage Induction Motors

1. Full Voltage Starting and Direct Coupling

The most common electric motor for driving belt conveyors is the AC squirrel cage induction motor. These motors are often started directly across the line. There are five standard NEMA designs of squirrel cage induction motors that are differentiated with respect to speed/torque characteristics. Two of these designs, NEMA B and NEMA C, are generally utilized on conveyor drives.

Typical speed/torque curves for NEMA B and NEMA C motors are shown in Figure 9-14. In each case the relationship of speed vs. torque is nearly linear as the motor approaches full load speed. However, the starting torque vs. speed relationship for these two motor designs are nonlinear and quite different. The NEMA C motor has a higher static or locked rotor torque and a higher minimum (pull–up) torque. The result is that a NEMA C motor will provide the conveyor drive more breakaway and acceleration torque than a like power motor of NEMA B design.

Figure 9-14 Typical Speed vs Torque Curves of NEMA Design B and C Squirrel Cage Motors

One of the drawbacks of starting a conveyor in this manner is that the belting, as well as all other drive components, will be subjected to the peak starting torque of the motor (breakdown torque) each time the conveyor is started. This maximum starting torque is generally at least 200 percent of full load torque and can be as high as 350 percent of full load torque.

This drive type is very common with lower power drive installations (i.e., less than 75 kW). The keys to applying this drive type successfully is to ensure that the drive will provide enough torque to breakaway and accelerate the conveyor and that belting of sufficient strength is selected so that the starting tension does not exceed 150 percent of the its kN/m rating.



With this method of starting a conveyor no speed or torque control is available in either the starting or running modes. The acceleration time is a function of the load on the motor (i.e., a lightly loaded conveyor will start faster than a heavier loaded conveyor).

9-5. STARTING TYPES


2. Reduced Voltage Starting and Direct Coupling

The output torque of the AC squirrel cage induction motor at any speed is a function of the square of the applied voltage. Therefore, one way to reduce the peak tension that the belting is subjected to during a start is to temporarily reduce the voltage to the motor. The dotted line in Figure 9-14 shows one possible result of applying a reduced voltage start to the depicted NEMA C motor.

There are several types of equipment that can produce a reduced voltage start. However, today most starters are of the SCR solid state design and are available from fractional kW to 1000 kW.

From a conveyor design aspect the objective of this starting technique is to reduce the starting torque enough to protect the belting and other drive components while providing enough torque to start the conveyor. A conveyor that has a predictable, steady load is most desirable for this type of starting. A conveyor that becomes overloaded can pose a problem as the drive may not provide adequate starting torque. Subsequent adjustments to the reduced voltage starter to cope with the overload condition may increase motor input torque beyond original design limits.

This type of starting can provide passive (without feedback logic) torque control in the starting mode. However, no speed or torque control is available during the running mode.

9-5. STARTING TYPES


3. Full Voltage Starting and Fixed Fill Fluid Coupling

Fixed fill fluid couplings are physically located between the electric motor output shaft and the input shaft of the reducer. When the electric motor starts the coupling from rest there is a relatively small amount of fluid in the coupling working chamber. In this state the coupling transmits very little torque. This allows the electric motor to temporarily come up to full speed virtually unloaded with the coupling at 100 percent slip (i.e., the coupling output shaft is at zero speed). The electric motor travels through the acceleration (locked rotor, pull-up and breakdown torque) portion of the speed/torque curve disconnected from the load. The fluid level in the working chamber then increases as fluid flows by gravity from a delay chamber to the working chamber until enough torque is generated to breakaway the conveyor and accelerate it to full speed. Because the motor is already in the nearly linear area of the speed/torque curve, the belting and drive components are generally not subjected to motor breakdown torque. The acceleration time of the conveyor can be in the range of 5 seconds to approximately 30 seconds with this type of drive. It should be noted that for a given fluid fill level the acceleration time is affected by load (i.e., a lightly loaded conveyor will start faster than a heavily loaded conveyor). Once the conveyor is up to speed the fluid coupling slip is generally in the 2 to 5 percent slip range.

This type of drive provides passive torque control in the starting mode and no speed or torque control in the running mode.

There are multiple types of fixed filled fluid couplings including those with delay and double delay chambers. The proper selection depends on the conveyor moving inertia and should be selected by a qualified manufacturer.

Figure 9-2- Fixed Fill Fluid Coupling Photos Courtesy Voith Turbo, York PA



9-5. STARTING TYPES


4. Full Voltage Starting and Variable Fill Fluid Coupling

The equipment utilized with this type of drive is very closely related to the Fixed Fill Fluid Coupling. The only functional difference is that the fluid level in the working chamber of the fluid coupling is variable and can be externally controlled. The control system increases the fluid level to increase output torque and, correspondingly, decreases the fluid level to decrease output torque.

Speed or torque sensors are often used with this type of drive to provide feedback logic to a control system that provides active control in both the starting and running modes. Fluid cooling equipment can be used with this equipment so that heat can be removed from the fluid during prolonged periods of high slip (i.e., 60-plus second starting ramps).

Typical types of varible fill fluid coupling include scoop tube (Figure 9-4) which uses a tube inserted into the working chamber to scoop excess fluid and the drain type (Figure 9-5) that is constantly draining and refilling to maintain desired fluid levels.

Figure 9-4 - Scoop Tube Fluid Coupling Photos Courtesy Voith Turbo, York PA



Figure 9-5 - Drain Type Fluid Coupling

9-5 STARTING TYPES


5. Full Voltage Starting and Hydroviscous Coupling

This type of drive includes a wet disc clutch that can be located between the motor and the high speed side of a gear reducer or between the low speed side of the reducer and a drive pulley. Some clutch units are internal to a reducer enclosure while others are external and shaft mounted.

The clutch pack is generally spring released and hydraulically engaged. In practice, the clutch pack is in its unengaged mode when the electric motor is started. This allows the motor to come up to speed disconnected from the load. Hydraulic pressure is then applied to the clutch pack and torque is subsequently transmitted through the clutch pack. The amount of hydraulic pressure applied to the clutch pack results in an infinite degree of clutch slip from 100 percent to 0 percent.

Controls with feedback logic are often used to provide active control in both starting and running modes. Fluid cooling equipment, as with the variable fill fluid couplings above, are used during periods of extended high slip.

Figure 9-6 - Gearbox with Hydroviscous Clutch Photos Courtesy of Dodge Division Rockwell Automation



Figure 9-7 - Clutch Stack

9-5. STARTING TYPES


6. Variable Frequency Control and Direct Coupling

The mechanical configuration of this drive type is identical to the drive in 9-5B1, above (AC Squirrel Cage Induction Motor with Full Voltage Starting and Direct Coupling). The performance difference of this drive is due to external modification of the applied AC waveform.

The speed of an AC squirrel cage induction motor is a function of the number of stator poles and the applied line frequency. A change in frequency will produce a proportional change in motor speed. For example, a four pole motor at



60 hertz will rotate at 1800 rpm, 1500 rpm at 50 hertz and 900 rpm at 30 hertz. In addition, motor torque is modified by applied voltage. This combination provides precise control of motor speed and torque. The electronic controller that synthesizes the waveform is generally called a variable frequency converter (VFC).

VFC drives are available for fractional kW through several thousand kW size motors.

This drive type provides active speed and torque control during both starting and running modes.

9-5. STARTING TYPES


7. Full Voltage Starting and Magnetic Coupling

This coupling works on the fundamental principal involving permanent magnets. The coupling consists of two separate components that have no physical contact. The couplings ability to transmit torque is created by the relative motion between the copper conductor and the magnets. This motion creates a magnetic field in the copper that interacts with the permanent magnets, thus transmitting torque through air.

Figure 9-8 - Magnetic Coupling

Photos Courtesy of Rexnord, Engineered Magnetic Technology Operation, Warren, PA

9-5. THE POWER UNIT

C. AC Wound Rotor Motor with Direct Couplings

For AC squirrel cage induction motors the shape of the speed/torque curve is an inherent characteristic of the motor. The wound rotor motor, however, permits control of torque, in steps, through the starting period by addition of external resistance to the secondary winding of the motor, which is electrically accessible through the slip rings. This permits a torque program during the start that can be planned to suit the particular conveyor the motor is to drive.

The general requirements of such a torque program are, first, a sufficient torque at standstill to breakaway the static forces, and second, sufficient torque to accelerate the loaded conveyor to running speed in a reasonable time. Furthermore, a period of time is required for starting to limit accelerating forces to values acceptable to the belt, the counterweight, vertical curves, or other critical considerations of the design.

The speed/torque curves presented in Figure 9-15 show the type of starting program used for a wound rotor motor driving a belt conveyor. Seven variations of secondary resistance, including the final, are shown. The points A, B, and C show how the rpm at which the motor absorbs full load is lowered by adding secondary resistance.



This type of drive generally provides passive (no feedback controls) torque control during the starting mode and no torque control during the running mode.

Figure 9-15 Wound Rotor Motor Speed Torque Program for Belt Conveyor Starting

9-5. THE POWER UNIT

D. DC Motor with Direct Couplings

The shunt or compound winding DC motor is generally used on belt conveyors. These motors provide an operating speed that is a linear function of the armature voltage. At zero armature voltage the output speed is zero. At 100 percent armature voltage the speed is 100 percent. These motors also provide constant torque from zero speed to 100 percent speed. This 100 percent or base speed can be raised by decreasing field current or lowered by increasing the field current. The limit to raising base speed is determined by the commutation limit. Once the commutation limit is reached, the base speed can still be increased but at the sacrifice of decreased torque. This adjustable base speed feature is useful for obtaining multiple conveyor speeds (i.e., full production speed, idle speed, creep speed, etc.). The DC motor and controller are thermally rated for continuous operation at all speeds.

The DC motor and drive are available in sizes ranging from fractional kW to thousands of kW.

This drive type produces precise torque control at all speeds. The result is active torque and speed control is available in both the starting and running modes.

9-5. THE POWER UNIT

E. Progressive Movements of Belt During Start

Measuring head and tail acceleration of various conveyors has shown the distance and time by which the starting of the head pulley precedes that of the tail. This varies with the nature of the belt and with the accelerating force.

The belt obviously absorbs energy during the starting period as it elongates under the accelerating forces. Then, when acceleration of the head begins to drop, this energy is given up by the belt and results momentarily in acceleration of the tail beyond its normal running speed as shown in Figure 9-16.

Figure 9-16 Tail Acceleration



In the ordinary conveyor, these relative displacements of head and tail are of little practical consequence, but they do make it obvious why sequence starting of multiple unit conveyors, rather than simultaneous starting, is necessary from a belt standpoint alone. They also show why relatively low-starting torque is sufficient to move even very long conveyors since static friction is overcome progressively from head to tail rather than simultaneously at all rollers. Even on steel cord belts, this progressive starting of the idlers is of substantial amount. On one 2400-m conveyor using steel cord belt, more than 6 m of belt passes the drive pulley before all the conveyor is in motion (see Figure 9-17).

Figure 9-17 Progressive Starting

Figure 9-18 illustrates these relationships for a steel cord belt with a relatively low acceleration. The drive is a slip ring motor with a motor-driven drum controller. The numbers on this curve indicate points at which the various contactor of the control close.

Figure 9-18 Acceleration Relationships

Figure 9-16 (above) illustrates a similar situation with a fabric belt with a greater acceleration. The drive in this case is a squirrel cage motor started on reduced voltage. A traction fluid coupling connects the motor and reducer.

Figure 9-19 shows the acceleration, at the head end only, of a fabric belt driven by a squirrel cage motor started directly on the line.

Figure 9-19 Head Acceleration



de 309-1 The Drive, Pulleys, and Motors: Introduction


TAKEUP

10-1. PURPOSE OF TAKEUP

The purpose of takeup devices in belt conveyors is to establish and preferably to maintain a predetermined tension in the belt at some critical point.

Often the critical point lies immediately following the drive since it is necessary to maintain tension at that point to prevent slippage on the drive pulleys. In some cases, however, the critical point may not be near the drive but at some point of lower elevation such as the foot of a slope or the bottom of a dip in the conveyor line. If so, the takeup must provide proper tension at such points even though doing so gives more tension than necessary at the drive (see Section 10-5).

"Takeup" devices derive their name from the fact that they take up changes in belt length. In taking up length, they maintain tension, which is their primary purpose.

10-2. TYPES OF TAKEUP

Takeups used with belt conveyors are of two general classes, manual and automatic. Manual takeups are usually screw operated and permit moving a pulley, usually the tail, to tighten the belt. These screw takeup devices give no indication of the tension they establish and are adjusted by trial methods until slippage is avoided. They are unable to compensate for any length changes in the belt between adjustments and so permit wide variations in belt tension. These imperfections are acceptable because of the simplicity of the screw takeup but only in short and lightly stressed conveyors where the penalties are not severe. Inclined conveyors up to 30 m centers and horizontal conveyors to 100 m centers would be in this classification.

Automatic takeup devices maintain a predetermined tension at the point of takeup regardless of length changes resulting from load change, stretch, etc. Usually, automatic takeup devices depend on gravity but are occasionally actuated by a torque motor or by hydraulic pressure.

Automatic takeup devices may be located at any point along the return run or at the tail pulley.

Tail pulley takeup is provided by mounting the tail pulley on a carriage that runs on tracks parallel to the centerline of the belt. A system of cable and sheaves to a suspended weight at any convenient location keeps this carriage pulled back and establishes belt tension. Takeup devices along the return run use a vertical or a horizontal festoon in the belt. With a vertical festoon, weights necessary to provide proper tension at that point are hung from the lower pulley of the festoon. A horizontal festoon requires cable and sheave connection to a hanging weight as in a tail pulley takeup.

Figures 10-1 through 10-4 show these three arrangements of automatic takeup and an arrangement for locating a counterweight on the slackside of a tail drive.

Figure 10-1 Tail Pulley Takeup

Figure 10-2 Vertical Festoon Takeup

Figure 10-3 Horizontal Festoon Takeup

Página 1 de 810-1 Takeup: Purpose of Takeup


Figure 10-4 Tail Drive, Takeup on Slackside

10-3. MOVEMENT REQUIRED OF TAKEUP

A. Manual Takeup

A manual takeup (usually a screw type) must provide enough movement to establish initial tension in the belt and to provide for periodic readjustment as the belt stretches. If the belt is pulled up with clamps before fastening so that it is at or near its running tension, the amount of movement used in establishing initial tension is reduced. Under such circumstances, a screw travel of 1 to 1-1/2 percent of the center distance is sufficient to give reasonable freedom from refastening problems. In the event that takeup travel is wholly used up, it is necessary only to cut out the old fastener and refasten at a shorter length. .


B. Automatic Takeup

An automatic takeup (counterweight most common) must provide sufficient travel to handle any elastic length changes due to load variation or to weather changes plus inelastic change (stretch) likely to occur between refastening or resplicing. In the case of fastened belts, it is not so important to ensure long periods between refastening; therefore, the amount of travel provided for stretch need not be so great. However, with vulcanized splices, years may elapse between resplicings and the takeup must provide for stretch during that period. With steel cord belts, it is occasionally necessary to increase takeup travel allowance to allow for thermal expansion or contraction where there are wide temperature changes.

In addition to these amounts that will be a percentage of belt length, an installation tolerance that is substantially a constant is required.

Takeup less than these recommended amounts does not mean that the conveyor will be inoperable but does mean that more frequent manual adjustments of length by resplicing or refastening will be required.

Occasionally it may be more convenient to increase center distance, usually by relocating the tail pulley, than to resplice the belt, or it may be desirable to splice at average loaded tension rather than empty tension.

Obviously, an automatic takeup must have sufficient travel to take care of elastic length changes in the belt because, without this minimum, it will strike the limits of its travel during every start and stop.



In cases where sufficient vertical height is not available to provide required movement for the counterweight, differential diameter cable drums can be used to multiply the movement of the weight. This requires a proportionate increase in the amount of counterweight.


C. Amount of Belt Stretch

The amount of elastic stretch that appears due to changes in load depends on the nature of the belt and on the degree of load change. Inelastic length change depends on the nature of the belt and the degree of belt stress.

Predicting the amounts of stretch with any degree of accuracy becomes difficult because of the many variables involved. A belt that would stretch 1 percent at one installation might stretch 1-1/2 percent at another. The following are some of the factors that can affect the amount of belt stretch in addition to the type of fiber, twist, crimp, and general belt fabric design.

1. Installation -- The tighter a belt is installed, the less stretch will be realized later.

2. Drive location -- Should a drive be located at the low tension end, the average belt tension will be substantially increased while maximum tension will increase little if at all. Total belt stretch will be increased because it is a function of average tension.

3. Moisture -- Moisture content can affect belt stretch and, if high enough, can even cause shrinkage.

4. Loaded starts -- Loaded starts impart a surge of tension to the belt above operating tensions. If such starts are frequent, it is obvious that the belt may stretch more than if started only occasionally.

5. Type of starts -- In addition to the frequency of starts, a motor that starts across the line will usually impart a much greater tension surge to the belt than a motor that has carefully con trolled start to limit torque.

6. Braking -- Belts that are stopped by brakes may be subject to overstresses and, as with starting, the greater the stress and frequency, the greater the stretch.

The elastic length changes in a belt due to load changes can be calculated with a degree of accuracy that hinges on the accuracy of tension calculations and elastic modulus. The following is an example of the calculations of estimated total takeup movement between empty running and full load acceleration:

Assuming average empty running tension is 3 kN/ply/m average loaded acceleration tension is 12 kN/ply/m (tensions calculated as in Section 6 and average tensions calculated as in Section 13), and elastic modulus (E = ∆ stress/ ∆ strain) equals 1500 kN/ply/m,

Generally, one percent will take care of elastic length changes in most fabric belts while one-tenth of one percent will suffice for steel cord.

10-4. INITIAL POSITION OF TAKEUP

The manner in which the belt is installed and fastened (or vulcanized) will determine the initial position of the takeup and influence the length of service before resplicing is necessary. In most cases, fabric belts should be installed so that substantially all the takeup is available for subsequent increases in belt length. An exception may be made with light loading or where the takeup is liberal in amount and a reserve of belt for possible repair or resplice can safely be stored in the takeup.



Table 10-A shows the recommended initial takeup position. This initial takeup position is that which the takeup finds after the clamps have been removed and the belt run empty a few belt revolutions to produce a natural tension distribution.

To install the belt so that the takeup will reach the desired initial position after starting and redistribution of tension, refer to the procedure outlined in Section 13.

Table 10-A Recommended Initial Takeup Position

Belt Type Percent available for length increase

Percent available for length decrease

Steel cord 1/4 Splice length

Fabric belt 90 10

10-5. EFFECT OF TAKEUP LOCATION

The takeup device can be placed at any convenient location along the return or at the idle terminal pulley. It is not necessary that it be placed adjacent to the drive, although this is often a convenient location.

Two primary considerations govern takeup location in most cases. First is the geometry of the conveyor. Frequently, the head end elevates to a loading pocket or tipple and provides enough height for a vertical takeup. In other cases, a ravine or ditch crossed by the conveyor provides height. Second is tension distribution around the conveyor. If analysis of the belt tension shows some location where tension is low, as at the foot of a slope, takeup can be accomplished there with a minimum of counterweight (methods of determining correct amount of counterweight for various takeup locations are discussed in Section 6).

If the takeup is located remotely from the drive, the acceleration of the drive and the amount of takeup counterweight must be related so that the counterweight will accelerate the portion of belt between the drive and takeup at the same or greater rate than the drive accelerates. Those cases in which remote takeups have failed were due to too rapid acceleration by the drive or too low acceleration by the takeup resulting in an accumulation of belt between drive and takeup.

In those cases where the takeup carriage is traveling other than horizontally, the effect of that component of carriage and pulley weight parallel to the takeup travel must be added or subtracted in determining the actual counterweight necessary.

Acceleration or deceleration forces sometimes affect amount of counterweight required. For a further discussion of the magnitude of such forces, see Section 6.

Table 10-B shows common conveyor arrangements and indicates those circumstances under which accelerating or braking forces are critical and cannot be ignored.

Table 10-B Effect of Acceleration or Braking on Counterweight Takeup

Conveyor geometry Preferred takeup location Acceleration effect Braking effect

Horizontal, head drive Following drive NoneTends to lift counterweight.

Brakes not usually large enough to cause trouble

Incline, head drive Following drive or at tail Little or none Little or none

Decline, tail drive At or near head Tends to lift counterweight if decline is slight None

Decline then level portion, tail drive At or near head

Critical - Lifts counterweight and feeds slack to foot of incline*

None

Critical when stopping with



*Such takeup problems can be handled by a very heavy, single counterweight, a double counterweight, by tail-end brakes, or head and tail driving.

Combinations of incline and decline, head drive

Following head or low point in return run Little or none

decline loaded - Lifts counterweight and slack runs to foot of decline*

Combinations of incline and decline, tail drive

Following head or low point in return run

Critical - Lifts counterweight and feeds slack to foot of decline

Little or none

10-6. EFFECT ON COUNTERWEIGHT OF REVERSING BELT DIRECTION OR REVERSING STRESS AT DRIVE

Either actual reversal of belt direction or reversal of stress at the drive changes the tension distribution in the belt. After the reversal, a point of low tension may become a point of high tension and the counterweight at such a point may be lifted. Reversal of stress without change in belt direction can occur in conveyors having combinations of incline and decline when a condition of inclines empty and declines loaded becomes reversed, i.e., inclines loaded and declines empty.

In such a circumstance, the first solution is to make the counterweight heavy enough to hold the takeup down even when high tension passes through it. This is at the expense of belt tension and has a limit. A second solution is provision of two takeups (Figure 10-5), one on each side of the drive, each weighted only enough to remain down when the belt at that point becomes the slackside of the drive. Under the reversed condition, one will be picked up but the other will become effective and take up the excess length.

With belt loaded full length, the conveyor is regenerative. Takeup B is then on the tight-side of the drive and is lifted. Takeup A is effective and controls tension.

With only the incline loaded, the conveyor requires power. Takeup A is then on the tight-side and is lifted. Takeup B is effective and controls tension.

Another method used on belts run in both directions is to have a single takeup located between two separate drives. One motor drives one direction; the second motor drives the opposite direction. In each case, the inoperative motor is free wheeling. The takeup tension must be enough for the higher of the two requirements.

Figure 10-5 Double Takeup Arrangement for Reversal of Tensions

10-7. RECOMMENDED MINIMUM TENSION

Due to conveyor geometry, it is entirely possible for the conveyor to have some points with little or no tension and still have sufficient slackside tension for driving. For example, a head-drive inclined conveyor may have a belt weight slope



tension sufficient to provide adequate slack side tension for driving and still have zero tension at the tail pulley. Since it is impractical to operate with zero or near zero tensions in the conveyor, certain minimum tensions for tail end or other low tension regions are provided. Table 10-C shows the minimum recommended values for various belt widths and material weights. Where the location of points of low tension is not obvious, the methods of Section 6 will disclose them.

Notes:

1. Tensions, loads, or idler spacings not shown can be directly interpolated.

2. This table is based on the following formula:

Table 10-C Minimum Recommended Belt Tension in Newtons (based on 2 % sag between idlers)

Combined weight of belt + load in kg/m (B + T/3.6S)

Idler spacing in metres at point of minimum belt tension, To

0.30 0.35 0.40 0.45 0.50 0.55 0.60 0.65 0.70 1.00 1.25

25 459 536 613 689 766 842 919 995 1 128 1 531 1 914

50 919 1 072 1 225 1 378 1 531 1 684 1 838 1 991 2 297 3 063 3 828

75 1378 1 608 1 838 2 067 2 297 2 527 2 756 2 986 3 445 4 594 5 742

100 1838 2 144 2 450 2 756 3 062 3 369 3 675 3 981 4 594 6 125 7 656

125 2297 2 680 3 063 3 445 3 828 4 211 4 594 4 977 5 742 7 656 9 570

150 2756 3 216 3 675 4 134 4 594 5 053 5 513 5 972 6 891 9 188 11 484

175 3216 3 752 4 288 4 823 5 359 5 895 6 431 6 967 8 039 10 719 13 398

200 3675 4 288 4 900 5 512 6 125 6 738 7 350 7 962 9 188 12 250 15 313

225 4135 4 824 5 513 6 201 6 890 7 580 8 269 8 958 10 336 13 781 17 226

250 4594 5 360 6 125 6 891 7 656 8 422 9 188 9 953 11 485 15 313 19 141

275 5053 5 895 6 738 7580 8 422 0 264 10 106 10 948 12 633 16 844 21 055

300 5513 6 431 7 350 8 269 9 187 10 106 11 025 11 944 13 781 18 375 22 969

325 5972 6 967 7 963 8 958 9 953 10 949 11 944 12 939 14 930 19 906 24 883

350 6432 7 503 8 575 9 647 19 718 11 791 12 863 13 934 16 078 21 438 26 797

375 6891 8 039 9 188 10 366 11 484 12 633 13 781 14 930 17 227 22 969 28 711

400 7350 8 575 9 800 11 025 12 250 13 475 14 700 15 925 18 375 24 500 30 625

425 7810 9 111 10 413 11 714 13 015 14 317 15 619 16 920 19 524 26 031 32 539

450 8269 9 647 11 025 12 403 13 781 15 160 16 538 17 915 20 672 27 563 34 453

475 8729 10 183 11 638 13 092 14 546 16 002 17 456 18 911 21 821 29 094 36 367

500 9188 10 719 12 250 13 781 15 312 16 844 18 375 19 906 22 969 30 625 38 281



where

To = minimum belt tension in newtons;

B = belt mass in kg/m;

T/3.6 S = load mass in kg/m;

where

T = metric tons per hour;

S = belt speed in m/s;

11 = carrying side idler spacing in metres; and

Sag = belt sag between idlers in metres(2 percent of 11 used in table).

10-8. AMOUNT OF TAKEUP TENSION REQUIRED

Previous portions of this section have discussed the various items affecting conveyor takeup. Table 10-D shows common conveyor arrangements with the formulas used to determine the counterweight tension. Where more than one formula is shown, solve both and use the larger. For arrangements not shown, see Section 6.

Table 10-D Amount of Takeup Tension Required, Arithmetic Calculation

Type conveyor Takeup location Takeup tension*

Horizontal, drive at head

Following drive

Table 10-Da.gif

Inclined, drive at head

Following drive

or Tail

Decline, drive at tail, belt is regenerative

Near head



*Counterweight required is twice the takeup tension plus or minus the effect of pulley and takeup carriage weight.

Te = Effective tension, Section 6

K = Drive factor, Section 9

To= Minimum tension at foot of decline, Table 10-C

gBH = Belt slope tension, Section 6

gCQL/2 = Return slope friction, Section 6

or Tail

Combinations of incline and decline Solve by methods of Section 6

10-9. MECHANICAL PRECAUTIONS AT TAKEUP

Horizontal or inclined takeup carriages must be assembled with a tight gauge between track and wheels to prevent misalignment. Hold-down rails should be used to prevent jumping the track in case of violent movement or dirt on the track. Vertical takeup carriages must be guided to maintain a horizontal axis on the pulley.

They cannot be allowed simply to hang in the belt festoon. A limit switch to open the drive control circuit should be provided near the end of the takeup travel to prevent operation without adequate takeup tension and to stop the conveyor in case of belt failure. Decking should be provided to prevent material falling between belt and takeup pulley, and guards should be installed to prevent maintenance personnel from being in the way of takeup travel in either direction.

Where takeup travel is minimal or where two takeups are used, there is a likelihood or a probability of the takeup carriage hitting its inner travel limit. Suitable rubber bumpers should be installed to prevent damage to either hardware or belt.

Section 14 also discusses takeup safety precautions.



CONVEYOR LOADING, IDLERS, AND AUXILIARY EQUIPMENT

11-1. LOADING

A. Feeders

1. General

The belt conveyor, moving rapidly, benefits by a continuous and uniform feed. Irregular or intermittent feed produces unloaded spots that result in decreased capacity or overloaded spots that result in spillage when excess material drops from the edges of the belt.

Ordinarily, the feed to the belt comes from some fairly large reserve of material; a continuous, uniform amount must be taken from this reserve to load the belt as long as it is running. With many materials, the belt itself cannot be depended on to do this job. Consequently, some mechanical feeding device is required.

Common types of feeders include the pan feeder, vibratory feeder, reciprocating feeder, belt feeder, and hand-controlled gates. These feeders are described in Items 2 through 6, below.

2. Pan Feeder

The pan feeder (Figure 11-1) is a widely used feeding device for heavy, lumpy materials such as ores and stone. It consists of a series of overlapping steel pans carried on chains and shears its load from the bottom of a hopper or pile. Its feed rate to the belt is controlled mainly by varying the speed at which its chains are driven and also by the front height of the opening through which the feeder drags its load. The speed of the feeder and its tonnage rate can be set manually by estimating the size of the resulting load on the belt or by referencing to an ammeter in the belt conveyor motor circuit that can be calibrated in tons per hour. Its rate also can be controlled automatically by using belt conveyor motor current as an indication for increased or decreased feed rate.

Figure 11-1 Pan (Apron) Feeders

With some materials, the feed from a pan feeder tends to be irregular since the material carried on each pan breaks separately and drops from the end to the belt in a series of pan-sized surges. Surges can be seen as variations in the load cross section on the belt. This effect is mitigated by heavy chains dangling against the discharge end of the feeder to break up the surges and is further smoothed out in the chute. Because of their size, pan feeders cannot be set down close to the belt; consequently, the material must fall several meters from feeder to belt. This fall sometimes makes an energy absorption problem, particularly if a chute cannot be interposed.

Página 1 de 4011-1A Conveyor Loading, Idlers & Auxiliary Equipment: Loading: Feeders

18-05-2007file://C:\WINDOWS\Temp\~hhC74D.htm

Pan feeders are preferably set parallel to the belt, and the belt is extended back under the feeder to collect any leakage of fine material through the pans. The pan feeders can be set horizontal, or they can be inclined above horizontal.

3. Vibratory Feeder

The vibratory feeder (Figure 11-2) consists of a suspended deck vibrated at high frequency and low amplitude by a vibrator motor. The motor is set so that the direction of the vibration throws particles of material upward and forward and so that they progress along the deck in a minute sawtooth path. The feed is controlled by the angle at which the deck is set and by the amplitude of the forced vibration.

Figure 11-2 Vibratory Feeder

The deck usually is set with a decline small enough that material will not slide without vibration. Using a slightly declined deck permits larger capacity from a given feeder. This type of feeder does not work well on some materials that are good energy absorbers and that, as a result, are not projected by the vibrations of the deck, but it does work well on a wide variety of materials. It also gives a very uniform feed and has the advantage that it can be set very close to the belt, which greatly reduces the impact of lumps.

4. Reciprocating Feeder

The reciprocating feeder (Figure 11-3) is related to the vibratory feeder, but its motion is one of high amplitude and low frequency. Its deck is moved in a reciprocating motion by a crank and connecting rod. On the forward stroke, deck and load are carried forward, and new material drops from the pocket or pile onto the back of the deck.

Figure 11-3 Reciprocating Feeder

On the return stroke, the material is kept from moving back, and the deck slides back under the load and drops material from its forward end. This is an intermittent feed, but the frequency is high enough and the flow is smooth enough in the chute to provide a good feed for such materials as coal. The deck sliding under the load on the return stroke is a handicap when the material is abrasive.

5. Belt Feeder

A belt feeder (Figure 11-4) is no more than a pan feeder of rubber and fabric suitable for less brutal service. The feed is controlled by varying belt speed and by the size of the opening through which the belt feeder drags its load. The sliding of the belt under the material as it drags its load out is destructive to the belt; thus, this type of feeder is not used for large lumps or sharp, heavy ma terials. The belt feeder is good for fine materials that might tend to leak away in mechanical feeders and for materials that will not feed on a vibratory feeder. Uniformity of feed is very good.



Figure 11-4 Belt Feeder

6. Hand-Controlled Gates

Free-flowing fine materials can be fed by opening gates until the proper rate is produced and a uniform flow is maintained. Other materials not fully free flowing are fed by hand-controlled gates (Figure 11-5) but with constant attention to gate opening. Gravel is recovered from stockpiles by this method, and the holds of self-unloading boats are discharged to the belts with a hand-controlled gravity feed.

Figure 11-5 Hand-Controlled Gate

11-1. LOADING

B. Chutes

1. General

Once the rate of feed is established, the direction, speed, and placement of material on the belt are controlled by the loading chute and skirtboards.

The receiving end of the loading chute is located, and its dimensions are determined by the feeder or the preceding conveyor. The chute's width must be great enough to accept material lying at the extreme edges of the preceding belt or feeder. Its elevation is determined by the trajectory of the material coming to it.

Material coming from feeders has nearly a vertical fall at all times; thus, the receiving end of the chute must be vertically below the feeder discharge but as high as possible without mechanical interference.

Because material coming from a preceding belt usually has substantial velocity, there is a choice of position for the receiving end of the chute. It can be placed near the horizontal centerline of the discharge pulley substantially tangent to the trajectory of the material. The advantage of this location is that it reduces impact on the chute and lessens material degradation. In this case, a dangerous condition exists during starting and stopping when the load has a trajectory too low to reach the chute. If the chute is extended back to the pulley to intercept a low trajectory, there is danger of lumps lodging between the chute end and the discharge pulley. This danger can be avoided by leaving a gap between pulley and chute through which the load drops for an instant during starting and stopping. This gap must provide passage for the largest lump (Figure 11-6).



Figure 11-6 Chute Profile for Normal Belt Transfer

The alternative position of the receiving end of the chute is low enough that it can extend back under the discharge pulley to receive material dropping vertically. The disadvantage of this position is that it prevents placing the chute anywhere near tangent to the normal trajectory and permits free fall of the material for some distance (Figure 11-7).

Figure 11-7 Chute Profile for Delivery from Feeder or Slow-Speed Belts

After the position of the receiving end of the chute is established, its slope is determined by the nature of the material, its entering velocity, and length and convergence of the chute. There is no attempt to give more than the roughest of rules for determining chute angle. To get the best flow through a chute, considerable experimental adjustment in the field is required.

Table 11-A gives chute angles commonly found for a few extremes of material types. If material enters the chute at more than 180 m/s from the preceding belt, the lower end of the angle range can be used. If the material has little entering velocity, the upper end of the angle range should be used.

Table 11-A Chute Inclination

Material Normal angle above horizontal (deg)

Sticky ores, clay, and earth 50 to 60*

Run of mine coal (damp) 35 to 45



*Water lubrication of chute sometimes improves flow. Bottomless chutes are sometimes used.

If there is substantial convergence of the chute, the angle of the intersection line of sides and bottom with the horizontal should be the governing angle rather than the angle at the centerline of the chute.

If any change of horizontal direction is required at the chute, the major portion of the chute should be in line with the receiving belt to achieve a central reloading. Also, loss of entering material speed should be allowed for in selecting the chute angle.

When rubber-lined chutes are used for abrasion resistance, it is more conservative to add 3 degrees to 5 degrees to the chute angle even though some materials slide as readily on rubber chute linings as they do on steel.

The discharge end of the chute must converge sufficiently to narrow the feed to approximately two-thirds the width of the receiving belt. This convergence should be uniform through the length of the chute.

It is desirable to let fine materials drop through to the receiving belt ahead of the lumps. This is achieved to some extent by notching the lower end of the chute back in a V shape or by using grizzly bars for the final 1300 to 600 mm of the chute. These grizzly bars are a hazard if there is any way they can trap the thin slabs coming down the chute on edge. Hence, they should have no cross members, and their opening should become wider toward the outer end. It also is beneficial to the belt to change the direction of the material toward horizontal at the lip of the chute although such upturned chutes are more apt to plug and to require more maintenance against abrasion.

2. Rock Boxes

Abrasive material sometimes is transferred from one belt to another through rock boxes. These boxes intercept the material trajectory with a partial box that retains a portion of the material. Succeeding material spills from the box and either falls directly to the next belt or slides down a short chute to the next belt. The major abrasion is taken within the material itself, and the direction and velocity of the spill to the following belt can be well controlled. The main disadvantage of rock boxes is that they somewhat degrade friable materials and are not self-cleaning when different grades or sizes of material are handled (Figure 11-8).

Figure 11-8 Rock Box Transfer

Sand (damp) 35 to 40

Primary crushed rock 35 to 40

Gravel 30 to 35

Screened rock 30 to 35

Ground feeds 35 to 40

Whole grains 27 to 35

Pulp logs 15



11-1. LOADING

C. Skirtboards

The chute, having delivered its load to the succeeding belt, has no further control of the material. Since no chute does its job perfectly, the load has not taken the exact direction, speed, or shape required on the receiving belt. Hence, there will be some turbulence of the load and rolling on the belt. To prevent material falling from the belt or loading too close to its edges, skirtboards are used to confine it for a few seconds until it settles.

Skirtboards are vertical or are inclined slightly outward at the top. They are commonly set approximately one-sixth of the belt width in from each edge of the belt (Figure 11-9). This one-sixth must be considered as a guideline only, as this figure wastes too much space (and carrying capacity) outside the load zone on wider belts, and on narrow belt leaves too little space outside the chute for effective edge sealing.

The solid structure of the skirtboard is never brought tightly against the belt, but is left with a clearance above the belt surface which then is closed with an elastomer sealing strip fastened to the side of the skirtboard. The clearance between skirtboards and belt should increase in the direction of belt travel to permit any trapped material to be freed. It is important that this skirtboard (and any liners installed) present a smooth line opening toward the exit end of the transfer point. Any jagged edge or sawtooth profile can serve as an entrapment point to capture material that will then abrade the surface of the moving belt.

Theoretically, the skirtboard should extend in the belt’s direction of travel to the point where the material is fully settled into the profile it is to maintain for the rest of its travel on the conveyor. Commonly, this distance is two to three meters from the end of the impact zone. But the specific length appropriate for a given installation depends on a number of factors, including belt speed and material velocity. In addition, the need to control air movement and provide an enclosure or plenum for dust collection may require that a greater length of covered skirtboard be provided.

The need to keep the belt down on the belt support structure without rising up into the steel skirtboard is one reason to avoid the use of the half-troughed pulley technique and provide adequate transition distances. To minimize the fluctuation in belt travel that can push the belt up into a steel edge, skirtboard should end above an idler, rather than between idlers.

Figure 11-9 Common Skirtboard Arrangements



11-1. LOADING

D. Skirtboard Sealing

The sealing system should be installed outside of the skirtboard, where it is removed and shielded from the severe downward and outward forces of loading material. In many cases it is useful to provide a sacrificial wear liner on the inside of the skirtboard, to serve as a dam to keep the weight of the material off the elastomer sealing strips.

There are a number of skirtboard sealing systems commercially available which provide a multiple layer sealing system to contain material and prevent spillage (Figures 11-10a and 11-10b, below). These sealing systems should allow for simple maintenance in compensation for wear from constant impingement of material and abrasion from belt motion.

Skirtboard sealing strips should not be fabricated from strips of old belting, since the fabric portions of used belting tends to trap abrasive particles and wear the belt under the skirts.

Figure 11-10a Skirtboard Sealing System

Figure 11-10b Skirtboard Sealing System

(Photos courtesy Martin Enginering)

11-1. LOADING



E. Geometry of Conveyor at Loading Point

Loading chutes can discharge to horizontal, inclined, or declined belts. If ample space is available, it is usually desirable to load on a level portion of the belt.

Loading on steep inclines has two major disadvantages. One, there will be a marked increase in load turbulence as the load accelerates to belt speed, which sometimes causes a substantially increased cover wear rate. Two, where lumps and impact are a problem, the impact on the belt is more severe because of the angle between the chute and the belt. However, many conveyors are successfully loaded on inclines up to 16 degrees.

Loading on a decline aggravates the turbulence at the end of the chute and consequent rolling of lumps down the decline. This aggravation requires longer skirt boards and gives more spillage. Run of mine coal and rock have been loaded on declines as great as 14 degrees, but ordinarily it is better to flatten the decline at the loading point to no more than half this angle. Changing the inclination for loading requires a vertical curve (see Section 12).

11-1. LOADING

F. Intermediate Loading

Stockpile reclaiming belt conveyors are loaded at intermediate points through traveling hoppers straddling the belt. In some cases, these hoppers are very elaborate and are equipped with feeders. In other hoppers, the feed is controlled only by the hopper opening. Such a loading method has the additional problem of shifting the hopper, but the belt loading still is governed by the principles described previously. Where loading is done at more than one fixed loading point, provision is made for retracting the skirtboards at those loading points not in use to avoid interference with the load already on the belt.

Loading additional material on top of a partial load already on the belt is not a common industrial practice but is widely done on coal mine belts. Belts can be damaged by the load already on the belt interfering with the skirtboards for the next loading point. Also, when there is no regulated feed, pouring additional load on spots already fully loaded can cause spillage.

Protection against this overloading is provided by a limit switch placed under the belt preceding the second loading point. A full load on the belt depresses the belt between idlers enough to open this limit switch and shuts down the feed at the second loading point until the fully loaded section has passed.

11-1. LOADING

G. Loading Point Impact

1. General

It is unavoidable that some of the energy from the falling material is absorbed by the belt and its supports at the loading point. There is just as much energy absorbed when sand is falling a given distance at a given tonnage rate as when the load is large pieces of ore or rock. Yet, the impact problem is entirely different. The difference is partially in the concentration of the impact on small areas of the belt, which happens with large sharp lumps, and partially in the internal energy absorption that takes place in such materials as sand. The impact of individual particles of sand, or other small-sized material, produces no measurable damage to belt cover or carcass. Individual impact of sharp lumps of ore or rock, however, produces serious cover or carcass damage unless the intensity of the pressure is reduced when the lump strikes the belt.

The first means of avoiding belt damage is to design the chute such that:

1. The drop to the belt and the chute angle are as low as possible.

2. Lumps land on a bed of material fines rather than directly on the belt.



The next step is to reduce the intensity of pressure between lumps and belt by permitting the belt to deflect under the impact. The belt has some capacity to deflect within itself and to relieve the force against it. However, even if the belt is all rubber (having no carcass whatever), it will not deflect readily because of its shape factor. That is, its thickness is so small with respect to width and length that any distortion in the thickness direction is highly restricted.

The solution is to permit the entire belt to deflect, which is what happens between supporting idlers when the load is placed on the belt at such a location. In such a case, the deflection of the belt is governed by idler spacing and by belt tension at the loading point. Absorbing energy in this way is limited because belt deflection becomes excessive, which results in spillage and in the trapping of material between belt and skirtboards where it can abrade the belt. When idlers are moved closer together to limit deflection, it becomes impossible to avoid loading lumps directly over the idlers. When the combination of belt tension and load necessitates that idlers under the loading point be closer than 36 in., this loading limitation is encountered and is a common condition.

While belts now can be designed to withstand rather substantial impact forces over steel idlers, the common practice today is to equip the loading point of virtually all bulk material conveyors with impact idlers (Figure 11-11) or impact cradles (Figure 11-12).

Figure 11-11 Grooved and Semi-pneumatic Impact Idler Rolls

Figure 11-12 Impact Cradle (Photos courtesy Martin Enginering)

To absorb the impact, idlers have been made readily deflectable. This has been attempted by supporting the entire idler on various springing arrangements, including rubber in shear, but the inertia of the idler is so great that this method has little value. A more effective method is to make the roll surface deflectable itself by using various designs of rubber rolls or semi-pneumatic tires for the roll. At one time, plain rubber-covered rolls were the only impact rolls available. Due to limited rubber thickness and lack of provision for flow of the rubber, these rolls had a limited capacity. Now, excellent energy-absorbing idler rolls of the general shape shown in Figure 11-11 are widely used. These rolls greatly increase capacity because the rubber depth and the grooving permit substantial deflection. They have energy-absorbing capacity several times that of the smooth rubber-covered rolls.

An alternative to the impact idlers is the use of impact beds or cradles (Figure 11-13). These cradles are installed under the loading zone to absorb impact and stabilize the belt path to minimize material spillage. The cradles are formed of multiple-layer impact bars with the top layer a low-friction material to allow the belt to ride over and a sponge-like secondary layer to absorb the energy of impact. These bars are mounted in a heavy-duty base to withstand punishing impact.

Figure 11-13 Impact Cradle (Photos courtesy Martin Enginering)



Figure 11-14, below, shows the relative energy that can be absorbed, from a single lump of sharp material, before the belt is damaged on steel idler rolls and on the grooved rubber type of impact roll and how the costly remedy of adding plies may be less effective than properly placing two or three impact idlers.

Figure 11-14 Belt Impact Energy Absorption Characteristics

2. Calculating Impact



In determining what degree of protection against loading irnpact is necessary, the following logical procedure should be observed:

First, determine what maximum lump size and weight will be.

Second, determine the distance lumps will fall from feeder or other source before they strike the belt. If no chute is interposed, the belt and its supports will have to absorb the entire energy of the fall (i.e., mass of lump (in kg) times distance dropped (in metres) gives joules of energy to be absorbed). This calculation quickly shows the importance of a good chute. For instance, a 50-kg lump falling 1.5 m would require 735 J of energy absorption.

Third, if a chute is used, the energy absorption is that resulting from the vertical component of the lump velocity at the lower end of the chute plus that resulting from the free fall at the end of the chute. This absorption will vary between the ideal condition, where the material velocity at the chute end (in the direction of belt travel) is equal to belt speed, and the extreme condition, where there is little or no friction loss in the chute and emerging velocity is substantially that resulting from a free fall. It is not safe, with heavy and sharp lumps, to assume that an ideal chute will be achieved and that only the energy resulting from such delivery to the belt must be absorbed.

The following method can be used to estimate energy to be absorbed under various loading circumstances:

Case I - for free fall to the belt (no chute)

I = gmh

where

I = energy to be absorbed by the belt and its supports in joules,

m = mass of lump in kilgrams

h = fall to the belt in metres

g = gravitational constant = 9.8 m/s²

Case II - very steep chute with little loss of gravity acceleration in chute (60 deg or more)

I = gmh (sin a)2

where

a = the angle between chute and horizontal.

Case III - normal chutes, 45 deg or less, where acceleration in chute is small

where


g = 9.8 m/s² (gravity acceleration)

h1= drop from end of chute to belt in metres.



Use, as an example, a material having 50-kg lumps corning from a feeder 1.5 m above the belt surface. If there is no chute and Case 1 is applied:

I = gmh = 9.8 x 50 x 1.5

= 735 J to be absorbed.

In Case II with a chute at 60 deg:

I = gmh(sin 60 deg)2

= 9.8 x 50 x 1.5 x 0.866²

= 551 J

In Case III with a 45-deg chute, a .25 m. drop from the lower end of the chute, and a belt speed of 2.5 m/s:

3. Belt Constructions for Impact Service

Selecting a proper belt construction for impact resistance is not a precise procedure that can be laid down in tables or formulas, except in a very general manner. The final selection usually is a matter of judgment and experience in evaluating all available data. The energy to be absorbed is but one factor, and even this factor can vary in severity with different materials. For example, a 50-kg lump of coal probably would damage a belt less than a 50-kg lump of ore because (1) the lump of coal is larger and its impact thus will be spread over a greater belt area and (2) it is more apt to crumble to some extent at impact, thereby alleviating some of the impact forces. Other factors to be considered are belt speeds; loading angle; lump shape (pointed, smooth, jagged); and any available histories of previous belt performances.

With the belts manufactured today, impact damage problems are rare compared with the days of cotton and rayon belts. Once a carcass selection has been made to satisfy tension and load support requirements, additional considerations for impact are seldom needed as long as reasonable consideration has been given lump size and loading area design.

4. Special Impact Belts

Figures 11-15 and 11-16 show two means of alleviating impact with special short endless belts in the loading area. These methods rarely are used today but were more popular when only cotton was available for belt carcasses and when far less impact resistance could be built into the belt.

Figure 11-15 Arrangement of Impact Belt at Loading Point



Figure 11-16 Arrangement of Pad Belt at Loading Point

5. Ross Feeder

As a means of retarding heavy lumps in the chute from the feeder or crusher and thus reducing impact to a manageable figure, the Ross feeder is used. This feeder consists of a group of very heavy endless chains hanging from a drum above the chute and lying against the load in the chute. These chains are driven slowly so that the chute does not plug but still places substantial restraint on the sliding lumps.

11-2. TRAJECTORY OF MATERIAL DISCHARGED OVER CONVEYOR HEAD PULLEY

The path of a material leaving a belt is important in determining the location of chutes or receiving hoppers. The trajectory of a material leaving a belt is determined by the following three relationships:

1. The centrifugal force, which acts against gravity and determines where the material leaves the belt.

2. The velocity of the material the instant it leaves the belt.

3. The force of gravity acting on the material the instant it leaves the belt and thereafter.

The equations covering these three relationships are:

where



D = distance in millemetres material falls when acted upon by gravity,

m = mass of material in kilograms per cubic metre or per lineal metre of belt


g = acceleration due to gravity = 9.8 m/s²

R = radius from the center of rotation to the center of load in metres, and

t = time in seconds.

Centrifugal force acts in a radial direction. The action of centrifugal force is opposed to the radial component of the material weight on the belt. This component can be calculated by the following equation:

Radial component of material weight = w cos A

where

m = material mass in same units as used in the centrifugal force equation (mS2/R)

A = angle measured from the vertical, either clockwise or counterclockwise as indicated in Figure 11-17, at which it is desired to calculate the radial component; in Case (2) of Figure 11-17, the radial component is equal on each side of the vertical as long as the angles are equal.

Figure 11-17 Belt Pulley Geometry for Centrifugal Discharge

Since the radial component of the material weight acts in the same path, but in opposite direction to the centrifugal force, the point at which they are equal determines where the material will leave the belt. Thus, the following relationship exists where the material begins to leave the belt.



When the value of S2/gR exceeds the cosine of the angle of incline or decline of the approaching conveyor, the material will start to leave the belt where the belt becomes tangent to the pulley because the centrifugal force exceeds the radial component of the material weight. There are six general cases showing where discharge begins (Figures 11-18 through 11-23).

After the discharge point is determined (Figures 11-18 through 11-23), a tangent is drawn to the pulley at that point. This tangent would represent the path of the material if the material were not acted on by gravity.

Figure 11-18 Low-Speed Level Belt

Figure 11-19 High-Speed Level Belt

Figure 11-20 Low-Speed Incline Belt

Figure 11-21 High-Speed Incline Belt

Figure 11-22 Low-Speed Decline Belt

Figure 11-23 High-Speed Decline Belt

The distance (x, mm) that the material would advance along the tangent equals 1000 St). However, since the material also



is acted on by gravity, it falls a vertical distance (D, in mm) equal to 500 gt2, using the same intervals of (t) as in calculating values of (x). The following two examples show how to calculate and plot the material trajectory.

Example 1 (Figure 11-24)

Assume:

Belt speed = 1.5 m/s,

Radius (R) = 0.5 m(pulley radius plus one-half load depth),

v = 300/60 = 5fps, and

Cos 18 deg = 0.951.

Therefore:

Since S2/gR is less than 0.951, the discharge point (Figure 11-20, above) will be at (a), where cos A equals 0.459 or A = 62.67 deg.

Using (t) intervals of 1/20 sec, then:

and



Figure 11-24 Trajectory of Material (Low-Speed Incline Belt)

Example 2 (Figure 11-25)

Assume:

Belt speed = 2.3 m/s,

Radius (R) = 0.5 m (pulley radius plus one-half load depth),

Cos 18 deg = 0.951.

Therefore:



Since S2/gR is greater than 0.951, the discharge point (Figure 11-21, above) will be at (b), where the belt comes tangent to the pulley.

Using (t) intervals of 1/20 sec, then:

and

Figure 11-25 Trajectory of Material (High-Speed Incline Belt)



The trajectory probably will deviate some if the material is acted upon by additional forces not considered in these trajectory calculations. For example, the adhesive force of cement to the belt may add to the total downward forces and the trajectory might be steeper than calculated; a very light material such as wood shavings may be acted upon by air resistance, which might elevate the trajectory. For the general run of materials, however, these forces are not present or are negligible, and the material will discharge from the belt tangent to the pulley and make a trajectory as calculated.

Using these two examples, there should be no difficulty plotting the theoretical trajectory of Figures 11-18 through 11-23, above.

11-3. IDLERS

A. Requirements Photos Courtesy Continental Conveyor & Equipment Co, Winfield, AL

Idlers are one of the most important components of a conveyor since they support the belt and load along the total conveying route. Although their function is only support, their importance is magnified by the fact that there are often several hundred or thousand individual idlers within the conveyor system. And each idler will often have several bearings and parts, meaning the reliability of the idler design is crucial to the overall reliability and availability of the conveyor.

Idlers are designed with rolls of varying diameter, length and numbers with antifriction bearings and seals mounted on shafts. The design of the rolls, bearings and seals make up a major part of the frictional resistance on a conveyor, which influences belt tensions and conveyor power. (Figure 11-25B)

The two main categories of idlers are carry (top) and return (bottom) as shown in Figure 11-25C. Carry idlers are configures with multiple rolls to form a trough. Typically, 3 rolls are used. However, 5 rolls are not uncommon. The return idler is normally needed just to carry and guide the belt back along the return of the conveyor. The most common return idler is made up of just 1 roll.

Figure 11-25B Idler Roll Cross Section

Figure 11-25C Typical Idler Configuration



Two roll "V" returns are sometimes used to facilitate belt training on long conveyors.

11-3. IDLERS

B. Classifications

The Conveyor Equipment Manuafacturer's Association (CEMA) has a classification system for the selection of the proper idler. Selection of the proper roll diameter, bearing size and shaft is based on the type of service, load and belt speed. Tables 11-B1- 11-B5 below show the CEMA ratings and the load ratings for each belt width and classification.

Each manufacturer has unique design criteria particularly for their bearing and seal design. For further information on specific recommendations, please reference the CEMA manual, Belt Conveyors for Bulk Materials, 5th Edition, Chapter 5 on contacting an idler manufacturer.

Table 11-B1

Classification Roll Diameter (mm) Belt Width (mm) Description

B4 100 450 through 1200Light Duty

B5 125 450 through 1200

C4 100 450 through 1500

Medium Duty

C5 125 450 through 1500

C6 150 600 through 1500

D5 125 600 through 1800

D6 150 600 through 1800

E6 150 900 through 2400Heavy Duty

E7 175 900 through 2400

Load Ratings CEMA B Idlers, kg (rigid Frame)

Ratings based on minimum L10 of 30,000 hrs at 500 rpm

Table 11-B2

Belt Width (mm)

Trough Angle Single Roll Return20o 35o 45o

450 186 186 186 100

600 186 186 186 86

750 186 186 186 75

900 186 186 180 70

1050 175 164 160 63

1200 172 160 155 60



Load Ratings CEMA C Idlers, kg (rigid Frame)


Table 11-B3

Belt Width (mm)

Trough AngleSingle Roll

Return

Two Roll

"Vee" Return20o 35o 45o

450 408 408 408 215 600 408 408 408 147 226

750 408 408 408 113 226

900 408 380 367 90 226

1050 385 358 347 68 226

1200 362 337 326 56 226

1350 340 316 306 226

1500 317 294 285 226

1600 226

Load Ratings CEMA D Idlers, kg (rigid Frame)


Table 11-B4

Belt Width (mm)


Return

Two Roll


450 544 544 544 272 600 544 544 544 272 750 544 544 544 272 385

1050 544 544 544 226 385

1200 544 544 544 419 385

1350 544 506 490 170 385

1500 521 485 469 127 385

1800 476 443 428 70 385

2000 385

Load Ratings CEMA E Idlers, kg (rigid Frame)


Table 11-B5

Belt Width (mm)


Return

Two Roll


900 816 816 816 453 590

1050 816 816 816 453 590

1200 816 816 816 453 590

1350 816 816 816 419 590

1500 816 816 816 385 590



1800 816 816 816 317 590

2100 816 759 734 249 590

2400 794 738 714 181 590

2500 590

11-3. IDLERS

C. Carrying Idlers Photos Courtesy of Continental Conveyor and Equipment Co, Winfield, AL

The most commonly used troughing idler has three rolls of equal length, with the inclined rolls usually at angles of 20 degrees, 35 degrees, or 45 degrees (Figure 11-25D). At one time, the 20-degree idler was standard in most applications, with the 35-degree and 45-degree idlers usually used only with grains and light materials. More recently, however, the higher angle idlers, especially those of 35 degrees, are being used much more widely in regular industrial applications.

The two primary reasons for using the higher angle idlers (35 degree and 45 degree) are to gain a greater carrying capacity and to gain more control over spillage, especially of lumps. The belt manufacturer generally encourages use of the lowest angle idler that will properly handle the desired load with minimal spillage. As the idler angle increases, more attention must be given to the following:

1. Belt constructions become heavier (and thus more costly) in order to support the added load properly and to avoid excessive creasing in the idler juncture.

2. Transition distances at the terminal pulley must be greater in order to maintain proper edge and center tensions (see Section 12).

3. Vertical curve radii also must be greater in order to maintain proper edge and center tension (see Section 12).

A number of special-purpose three-roll idlers are manufactured. A few of these idlers and their use are described below:

1. Picking idlers have short 20-degree troughing rolls and a long center roll and are used in picking, sorting, and feeding (Figure 11-25E).

2. Long center roll idlers are similar to picking idlers but have trough angles of 35 degrees and 45 degrees. They provide less capacity than the equal roll idlers, but the high trough angles help control spillage. They have been widely used to haul grain, wood chips, and other light materials.

3. Three-roll offset idlers (Figure 11-25F) essentially do the same job as the three equal-roll type. The difference is that all three rolls are longer, and the center roll is mounted out of the line and ahead of the two trough rolls so that the roll lengths overlap; this overlap is beneficial to the belt in that the idler

Figure 11-25D Inline Rolls

Figure 11-25E Picking Idler

Figure 11-25F Offset Rolls

Figure 11-25G Impact Idler



gap is eliminated, and belt contact points are much farther from the idler ends, thereby making it more difficult for excess idler grease to damage the belt cover.

4. Impact idlers of various designs are used to absorb abuse at the conveyor loading point (Figure 11-25G).

5. One- and/or two-directional training idlers of various designs are used to help steer the belt (Figure 11-25H).

6. Special transition idlers are available to change the belt cross section properly from troughed to flat at the conveyor terminals (Figure 11-25I).

Another idler that is finding more and more use is the flexible or catenary type (Figure 11-25J). Some of these idlers are made with molded rolls mounted on a flexible steel cord to form the catenary. Another type is made with several regular idler rolls connected at their ends with a flexible link. Still another resembles a coil spring, which provides impact cushioning and also can assume whatever trough angle the weight of the load may demand.

Figure 11-25H Training Idler

Figure 11-25I Transition Idler

Figure 11-25J Catenary Type

11-3. IDLERS

D. Return Idlers Photos Courtesy of Continental Conveyor and Equipment Co, Winfield, AL

It is not always recognized that return idlers are subjected to service conditions equal to or greater in severity than those of the carrying idlers;

Figure 11-26A Flat Returns



this is due to the greater spacing, greater span between bearings, greater bearing loads, and operation against the dirty side of the belt. Consequently, return idlers should be equal in diameter to carrying rollers. The length of the return rolls is particularly important since most edge damage to a belt is caused by the belt rubbing the conveyor structure along the return. At least the same roll length and clearance should be provided on the return idlers as for the pulley face width (see Section 9).

Rubber disc return rolls, which support the return belt on a group of rubber or rubber-faced discs, are used where sticky material carried by the belt tends to build up on conventional steel tube rolls. The concentration of pressure on the discs and their deflection tend to prevent material from caking. Rubber disc rolls should not be used with belts that will not lie flat laterally on the return run. Belts that tend to remain concave toward the top cover force their edges down between the discs and carry their entire weight on the discs near the edges, thus wearing them rapidly. This tendency to remain concave toward the carrying side is found in some fabric belts, particularly with fabrics woven specially for extra transverse flexibility, in some cord belts, and, to some small extent, in any belt with light carcass and heavy top cover. Where this condition exists, a return roller consisting of spaced discs in the center portion and a rubber sleeve at the roll ends is effective.

Helical or spiral return idlers as well as cage-type idlers with transverse slats also are available in various designs to minimize material buildup.

Figure 11-26B "V" Returns

11-3. IDLERS

E. Idler Spacing

For the ordinary conveyor, it is customary to place idlers at uniform spacing both on the carrying and the return run. Such spacings are a compromise based on averages of tension and of material weight for various belt widths. This practice results in greater sag at low tension points and less sag at high tension points.

Table 11-B, below, gives typical uniform idler spacings that are commonly used, but the idler manufacturer should be consulted for complete recommendations.

On long conveyors where considerable tension is developed to minimize sag, there is a growing trend toward adopting idler spacings considerably in excess of those in Table 11-B and as great as 3 to 4 metres. In considering these spacings, some precautions should be observed:

1. The idler manufacturer should be consulted to ensure that his design limits are not exceeded.

2. Belt sag should be carefully investigated not only at full-load belt tensions but also at empty belt tensions to guard against the possibility of excess spillage when only a very short section of the belt is loaded. Sag should be limited to a maximum of one to two percent when belt is loaded.

3. Belts can be lifted from idlers by high winds; greater idler spacings may increase this possibility.

4. There may be a greater tendency for large lumps to spill over edges with wide idler spacing due to edge sagging.



Table 11-B Typical Spacing Recommendations for Conveyor Idlers

Belt Width (mm) Carrying idler spacings (m)

Return idler spacings (m)Material weight (kg/m3)

Up to 800 801 to 1600 1601 or more

350, 400, 450 1.6 1.5 1.5 3

600, 750 1.6 1.2 1 3

900, 1050 1.5 1 .9 3

1200 1.4 1 .8 3

1350 1.4 .9 .8 3

1500 1.4 .9 .8 2.5 to 3

1800 1.4 .9 .8 2.5 to 3

11-3. IDLERS

F. Graduated Idler Spacing

Because of their length, some belts of great strength and wide tension range make graduated idler spacing more worthwhile. This practice makes possible a saving in the number of idlers required, compared with the average spacing practices, and results in better support of the belt in low-tension regions. Idler spacing, along with several other factors, is an important influence on conveyor friction.

Considerable experimental testing has been performed to determine the effect of the type of idler lubricant, idler load, idler spacing, temperature, and belt tension on the friction factor (C) in a conveyor system.

Figure 11-28 shows the effect of different idler lubricants and load conditions on the internal idler coefficient of friction. This figure shows a relatively constant coefficient for various loads except that, as loads become light, the constant drag items (such as bearing seals) assume greater importance and increase the coefficient of friction. The actual friction drag in ounces at the roll rim decreases continually with load. New grease makes internal friction slightly higher than grease in a well-runing idler.

The internal idler coefficient of friction of the running conveyor is shown by Figure 11-28 to be from 20 percent to 30 percent of the composite friction factor (C) used in the design of well-maintained conveyors with antifriction idlers.

Figure 11-28 Internal Idler Coefficient of Friction



NOTES:

Internal friction curves for idler rolls at different grease and load conditions

Constant temperature of 24 deg C

Used 125-mm OD idler roll with tapered roller bearing

The starting coefficient of friction increases sharply with heavier idler loads. However, this factor is not serious or objectionable due to the nature of the belt itself. Because of the belt's inherent elasticity, all idlers are not required to start at once, which reduces the overall effect; this is similar to a railroad train where all cars are not started simultaneously but consecutively. This action has been shown to be true even in steel cord belting where the elastic elongation of the belt during starting, which produces the consecutive starting effect, is very small.

Figure 11-29 shows the effect of temperature, idler lubricant, and load on the internal idler friction. In the normal ambient temperature range, where conveyors usually are run, there is little effect due to temperature. Where the temperature drops below 0 degrees C, the increased friction can be offset by a low-temperature bearing lubricant.

Figure 11-29 Internal Idler Friction Using Different Lubricants and Temperatures



Additional studies have been made on conveyors running at temperatures as low as -35 degrees F to -40 degrees C. Results indicate that, with proper greases, the friction increase above that at normal temperatures still is of little significance except during startup, when the difference can amount to 20 to 30 percent or more for about 15 minutes to one hour. It was concluded that the normal friction factors (Section 6) are adequate for belt tension calculations and selection in cold weather considerations except where frequent starts are anticipated or for prolonged operation at -40 degrees C or below. However, the drive motor selection should be referred to the motor manufacturer to ensure that startup overloads can be safely handled.

Figures 11-30, 11-31, and 11-32 present the same data. In each case, the ordinate represents the composite friction factor (C). Belt tension, load per meter of belt, and idler spacing are plotted on the abscissas, respectively. The internal friction of the idler itself is a relatively small proportion of the total friction. The remainder of the friction is due to the belt rolling on the idler and to the internal friction of the load on the belt. These curves show that, where the belt tension is very high, there is little internal friction of the material. With normal tension and idler spacing, the rolling friction is approximately 30 percent of the total, which leaves from 50 to 60 percent of the total as the internal friction of the load.

Figure 11-30 Idler Friction Test (Idler Spacing, Belt Load Curves)



Figure 11-31 - Idler Friction Test (Belt Tension, Idler Spacing Curves)



Figure 11-32 Idler Friction Test (Belt Tension, Belt Load Curves)



The proper relating of belt tension and idler spacing to obtain a uniform friction factor throughout the entire length of a conveyor installation cannot always be realized because the idlers cannot practically be placed close enough together in the low-tension area to achieve a uniformly low-friction factor. A uniform friction factor at a higher level possibly could be achieved by spreading the idlers in the high-tension end of the unit. However, this would increase the overall friction factor of the unit, a condition that is not desirable.

11-3. IDLERS

G. Diameter of Idler Rolls

Idlers are commonly made with roll diameters from 100 mm to 200 mm, with shafts, shells, bearings, and seals all designed for heavier use as diameters increase. Diameter selection is influenced by speed; weight of load; and belt, lump size, and life expectancy. The idler manufacturers should be consulted as to proper selections, but in general the 100-mm to 125- mm diameters are suitable for 1.5 to 3 mps belt speeds and for hauling materials up to 100 pcf. The 150-mm to 200-mm diameters can handle heavier materials (more than 100 pcf) at 4 to 7.5 mps.

Continental Conveyor and Equipment CoHewitt-Robins



GoodmanFMCRexnordPrecision

11-4. SELF-ALIGNING METHODS

A. General

Various means have been devised for attaining an automatic centering influence on the belt as it travels over its carriers. As discussed under "Pulley Crown" in Section 9, the terminal pulleys have very little influence on guiding the main portion of the belt. It is primarily the idlers or carrying rolls that control the centering.

On the top run, an automatic centering effect is obtained from the regular troughing idlers by tipping them forward in the direction of belt travel. A tilt of no more than two degrees off vertical on every troughing roll produces a strong self-aligning effect. This effect occurs because a tilt forward on a troughing idler throws the outer end of the inclined roll ahead of the inner end. When the belt tends to run to one side, an increasing portion of its weight rests on that roll, which is so canted that it tends to return the belt to the center (Figure 11-33).

Figure 11-33 Effect of Tilted Troughing Idler

The belt manufacturer recommends minimal use of tilted idlers (especially with high-angle idlers) since they can cause accelerated pulley cover wear due to extra friction between the belt and troughing rolls.

Special self-aligning troughing idlers are commonly used in place of, or in addition to, the tilting idlers to provide a self-aligning influence. These self-aligning idlers are pivoted about an axis vertically perpendicular to the centerline of the belt.



When the belt becomes off center, they swing about this pivot so that the axes of the rolls themselves become canted in a corrective direction. The self-aligning idlers swing about the center pivot in various ways usually associated with the pressure of the off-center belt against an eccentric disc at the end of the roll, a fixed arm attached to the idler frame, or a brake actuating arm; all tend to carry the end of the idler, to which the belt has run, in the direction of belt travel.

Self-aligning return rolls operating on the center pivot method also are used for automatically controlling the centering of the belt's return run. In some return idlers, two rolls form a 10-degree to 20-degree V trough, which is effective in helping to train the return run.

In some cases, special aligning devices have been provided on the return run just ahead of the tail pulley to ensure that the belt is brought centrally around the tail pulley and thus is loaded centrally. These devices usually consist of a group of three self-aligning return rolls connected with light cables so that they swing about their pivots in unison. In addition, they are provided with mechanical help in swinging so that they give a more prompt and greater corrective effect.

In some cases, this mechanical help in swinging comes from the edge of the belt at a point preceding the self-aligning idlers. This mechanical help is accomplished by a roller, bearing lightly against the belt edge, that transmits lateral belt movement through a bell crank arrangement and small cables to the pivoted idlers. Another way of doing this is to take the indication from the belt edges with light relays that open and close the circuit of a small motor, which then swings the pivoted idlers in a corrective direction. Figure 11-34 shows the general arrangement of such a device.

Figure 11-34 Return Belt Centering Device



de 4011-1A Conveyor Loading, Idlers & Auxiliary Equipment: Loading: Feeders



B. Aligning Problems in Reversing Conveyors

Titled idlers and some types of self-aligning idlers for training the belt cannot be used when the belt is reversed because what was a corrective effect in the forward direction becomes a disruptive effect in reverse.

Consequently, if a belt is to be reversed, the idlers must not be tilted, and the self-aligners must be able to swing about their pivot in a corrective direction regardless of belt direction. Those types that depend on friction of the off-center belt to shift the idler will work in both directions of belt movement. Those that depend on the pressure of the belt edge against an arm extended in front of the idler will work in one direction only.

Even with proper self-aligning idlers, training a reversing belt requires that all idlers and pulleys be aligned very carefully and that the conveyor structure be leveled and aligned since any imperfections cannot be corrected for by canting a few idlers, as in a one-way operation. Reversing long belt conveyors should not be lightly considered for these reasons and also for those reasons given in Section 10 relative to takeup problems.


C. Spacing of Self-Aligning Idlers

If everything about the conveyor is aligned and leveled and if the belt is troughable and straight, no self-aligning idlers are needed. Rarely are these imperfections wholly absent, however, and thus most conveyors need self-aligning idlers. The troughing idlers often will provide adequate training on the carrying side, but all conveyors should have self-aligners on the return, especially in the area preceding the tail pulley.

Self-aligners are needed under the conditions given in Table 11-C.

Table 11-C Conditions Requiring Self-Aligning Idlers

Top Run Return run Spacing

For reversing belts Same as top run 30 to 50 m apart, at least one idler on short center conveyors

For belts that do not trough well empty Same as top run 30 to 50 m apart, at least one idler on

short center conveyors

For belts inherently unstable Same as top run 30 to 50 m apart, at least one idler on short center conveyors

- - - For any belt at point just ahead of the

tail pulley to ensure centering under the loading point


D. Guiding Devices on Belt

Some conveyors impossible to train otherwise are guided by a rib, molded centrally along the underside of the belt that runs loosely in a groove in the idlers and pulleys. This guide is not a practical or a necessary means on long conveyors that can be trained otherwise. It is a last resort where out-of-level operation and off-center loading defeat ordinary training methods and for extremely short, wide conveyors that do not respond to other means of training.

11-5. CLEANING DEVICES

A. General



Cleaning adhered material from the belt surface before the belt enters the return run of the conveyor is important. Failure to clean the belt allows material to transfer from the carrying surface to idler rolls and snubs and to fall under the conveyor. Material buildup on the rolls and snubs will degrade belt training and increase drive power consumption. Fugitive material released into the plant environment can become a fire and safety hazard, as well as a significant maintenance and housekeeping expense.

In some cases, steps can be taken to mitigate the effects of a dirty belt. Dirt can be kept from building up on the snub pulleys by using soft rubber coverings or by directly cleaning a pulley. Rubber disc return rolls prevent dirt from building up on the rolls and, thus, may prevent training problems. But a more effective solution is the installation and maintenance of an engineered cleaning system.


B. Belt Cleaning Systems

While some operations continue to prefer "home-made" belt cleaning systems, the desire for clean belts and reduced fugitive material has led to the design of engineered belt cleaning systems by a number of specialty suppliers.

Typically, these belt cleaning systems consist of urethane or metal scraper blades, secured on a mainframe or shaft and positioned across the carrying width of the belt near the conveyor discharge. The scraper can be one continuous blade or a group of smaller, modular blades assembled to form a unitary or overlapping cleaning edge. The system is completed by a tensioner which presses the cleaning edge against the belt to remove material yet allows the edge to release away from the belt in the event of the passage of an obstruction, such as a non-recessed or damaged mechanical belt splice.

It is important that these belt cleaning devices be designed to minimize any risk that they could damage the belt. Cleaners will see significant abrasive wear, and so must be designed to allow efficient maintenance, including the retensioning of the cleaner against the belt and the replacement of worn blades.

The characteristics of the material have a significant bearing on the amount of carryback adhering to the belt and, therefore, on the efficiency of the belt cleaning systems. It is important to provide sufficient cleaning capabilities to perform under "worst case" conditions.

The choice between elastomer (typically urethane or rubber) and metal blades often depends on plant preference. Some operations see improved cleaning with metal blades; others prefer urethane to avoid putting a metal edge close to a valuable belt.


C. Cleaner Location (Photos courtesy Martin Enginering)

Cleaning should take place as far forward—as close to the discharge point—as possible (Figure 11-35 and Figure 11-36). To reduce the potential for material release into the plant environment and to provide a firmly supported surface for the cleaner to scrape against, the first cleaner (or pre-cleaner) should be installed on the face of the head pulley just below the trajectory of the discharging material. A secondary cleaner should be installed to clean the belt while it is still against the head pulley. If additional secondary cleaners are desired, or if the preferred position is unavailable due to space limitations, the cleaner(s) should be mounted to clean against snub pulleys, return rollers, or other components that give a firm profile to the belt.

If a multiple cleaner system is used, the pre-cleaners installed on the pulley face can remove the bulk of carryback. This cleaner should use elastomer blades applied at light--approximately 2 psi (14 N/m²)--pressure. This force reduces the risk of the damage to the blade or belt should a splice catch the cleaning edge, improving blade life and reducing belt wear.

Secondary cleaners incorporating metal or urethane blades can perform the final, precision or mop-up cleaning using somewhat higher blade-to-belt pressure. Research has indicated that the optimum blade-to-belt pressure is 11 to 14 psi (76 to 97 N/m²). Increasing pressure beyond this range raises blade-to-belt friction, shortening blade life, increasing belt wear, and increasing power consumption without providing any improvement in cleaning performance. Like shaving, it is more effective to use multiple passes with moderate pressure (i.e., multiple cleaners) than one high-pressure stroke over the



surface. Figure 11-35 Multiple Cleaning System

Figure 11-36 Multiple Cleaning System


D. Other Cleaning Systems

Rotating brushes have been used successfully in some belt cleaning applications, particularly on chevron, ribbed, or cleated belts (Figure 11-37). These powered brushes are most effective with dry materials. With wet or sticky materials—the material most likely to adhere to the belt in the first place--the bristles are apt to become clogged with residual material.

Where water is available and where waste disposal or freezing is not a problem, washing the belt with a water spray followed by wiping with one or more scrapers can provide a good cleaning job on almost any material.

Figure 11-37 Rotary Brush Cleaner (Photos courtesy Martin Enginering)


E. Cleaning Pulley Side of Belt (Photos courtesy Martin Enginering)



The only cleaning required on the pulley side of the belt is the removal of materials--principally lumps--that have fallen or bounced onto the return run. If not removed, these lumps can be carried into the tail pulley where they can damage the belt’s interior surface and the pulley, or build up on the pulley to promote belt slippage.

A rubber- or urethane-faced plow installed immediately in front of the tail pulley will deflect this material off the belt. Typically, this plow is v-plow for single direction conveyors (Figure 11-38 ) or a diagonal plow for reversing belts (Figure 11-39). These plows are tensioned against the belt by gravity. They should be suspended from points above and in front of the attachment to the plow, so that the impact of lumps does not drive the plow downward against the belt.

Figure 11-38 V-Plow

Figure 11-39 Diagonal Plow


F. Turnover

A different technology avoids the need for cleaning the belt rather than providing a means for doing the cleaning. This method involves turnover of the belt 180 degrees after the belt passes the discharge pulley. Thus, the pulley or clean surface of the belt is in contact with the return idlers. A similar turnover must be made at the opposite end of the conveyor to bring the conveying cover up again at the tail pulley (Figure 11-40).

Figure 11-40 Turnover Belt Drive

Brakes are essential on downhill belts to stop the loaded belt in case of power failure or other emergency. They also are necessary on any conveyor where excessive coasting after power interruption is apt to deliver materials to points where it cannot be handled.

While this method is finding more and more use, it is recommended from a belt standpoint that turnover be limited to those cases where sticky material cannot be adequately cleaned by more conventional methods. The turnover increases tensions at the belt edges and reduces center tension. This mal-distribution of tension can lead to belt buckling and even longitudinal foldovers if the turnover length is inadequate and/or other features of the turnover and its installation are not proper. The belt itself may have a natural tendency to cup slightly due to cover shrinkage or, even if flat when new, it may tend to cup later for a variety of reasons, which can lead to turnpver problems.

Section 12 describes more completely turnover design and installation methods.

Martin Engineering



Arch Environmental- GordonFlexco

11-6. COVERING THE CONVEYOR

Permanently installed conveyors usually are enclosed in a gallery of sufficient width so that maintenance and inspection personnel can pass readily along one side. Such a gallery protects personnel, machinery, and material handled from the weather. In addition, attention given a conveyor by workers exposed to rain and snow is not apt to be painstaking.

Where the conveyor is operated in moderate climates and where material is not affected by rain, operation with only a walkway beside the belt is feasible. If the material itself must be protected, partial galleries consisting of a wall on the windward side and a roof or a semicircular cover of corrugated sheet metal are commonly used. Where wet sloppy materials must be handled in cold weather, it is almost essential to provide a completely enclosed gallery and sufficient heat to prevent the material from freezing.

For the protection of the belt itself from the sun and rain, a high-cost gallery is hardly justified. Unless the climate is favorable, permanent conveyors usually require a gallery to protect the belt from strong winds and to protect men and material. For instance, heavy snow must be kept out of aggregate on conveyors since the snow will plug the chute.

11-7. CONTINUOUS WEIGHING

For most purposes, continuous weighing of material passing a given point on a belt is sufficiently accurate. The material is weighted by suspending a section of the conveyor from a beam balance. A continuous summation of the net weight is made by an integrating device driven from the belt itself. Such a continuous weighing device must be placed so that no vertical component of the belt tension can influence the beam position, since belt tension varies and cannot be compensated for in the tare weight.

11-8. MAGNETIC SEPARATION

Separation of magnetic from nonmagnetic materials on a belt is common. The simplest case is the removal of tramp iron (such as rail spikes and picks) from coal or ore by an electromagnet or permanent magnet suspended above the trajectory of the material at discharge. The material must be removed manually from the magnet.

Magnetic pulleys, both electric and permanent, are used in two ways. Most common is the magnetic head pulley, which holds magnetic objects or particles against the belt beyond the normal point of discharge, thus forcing them to fall into a separate chute. A second method uses a separate belt and magnetic pulley above the load on the main belt. Magnetic pieces are lifted from the load by the magnetic pulley. They then are carried up and around the pulley and over the separator belt to a disposal bin (Figure 11-41).

Figure 11-41 Three Alternate Magnetic Separation Methods



11-9. BRAKING AND RESTRAINING DEVICES

A. Brakes (Table 11-D)

Brakes are essential on downhill belts to stop the loaded belt in case of power failure or other emergency. They also are necessary on any conveyor where excessive coasting after power interruption is apt to deliver materials to points where it cannot be handled.

Conveyor brakes usually are kept released by application of power and are set by spring pressure or weight when the power is interrupted. Two common types are Thrustor and solenoid released band brakes that work on a brake drum on the high-speed side of the reducer. Used in this location, they can be relatively small brakes since they have a mechanical advantage through the reducer. There is the objection that mechanical failure between the brake and the drive makes the brake ineffective. Consequently, on downhill conveyors, a band brake, directly on the extended driving pulley shaft, often is used. In this location, the brake drum must be much larger, but its operation essentially is the same. The total decelerating pull at the braked pulley is converted to torque at the brake location whether the pull is on the pulley shaft or on the high-speed side of the reducer. Brakes are rated on a torque basis, and size now can be selected. Brakes for downhill belts should be liberal. Caliper or disc-type brakes are used on conveyors with considerable success.

Runaways on downhill belts are very seldom caused by mechanical failure between the brake and drive pulley. They occasionally result from the brake itself failing, due to lack of adjustment or inadequate heat capacity, or from the belt slipping on the drive pulley.


B. Anti-Rollback Devices (Table 11-D)

Inclined conveyors with sufficient grade tend to run backward when power is interrupted with the belt loaded. This backing up tendency is restrained by various anti-rollback devices. The ratchet and pawl arrangement on the head shaft is quite common, with various devices for holding the pawl disengaged as long as the conveyor has power or is up to speed. On large ratchets, the pawl is held up by a solenoid or Thrustor as long as the conveyor has power but drops on power interruption. In other cases, the pawls are in a rotating frame with the ratchet stationary and are held disengaged by centrifugal force until speed drops. These avoid clattering of the pawls on the ratchet during normal operation. When the Thrustor lifts the pawl, the Thrustor should be wired into the motor control circuit so that it does not lift until the starting program is well under way. Lifting the pawl with the first control contactor has permitted belts to back down the slope due to lack of motor torque on the first contactor, which was intended only to accelerate the empty belt.



In addition to ratchets, it is common to use self-energized differential band brakes either on the drive pulley shaft or on the high-speed side of the reducer. One-way clutches at the same locations as the band brakes also serve the same purpose. On very high slopes, it is common to use two anti-rollback devices; one is a band brake, one-way clutch, or ratchet on the head shaft and the other is a brake or one-way clutch on the high-speed side of the reducer.

Combinations of incline and decline in a single conveyor require investigation of various loading conditions to determine need of brake or anti-rollback. The requirement may vary from both brake and anti-rollback to neither. The combination of anti-rollback at the head end and brake at the tail is sometimes required. A brake at the tad avoids passing decelerating forces through the takeup and hence permits smaller counterweight in some cases.

Table 11-D Brake and Anti-Rollback Recommendations

Item Level Conveyor Incline Conveyor Decline Conveyor

Brakes

Need Only if coasting is objectionable Desirable on grades > 3% to supplement anti-rollback Essential

Type Thrustor or Solenoid on high-speed side of reducer

Thrustor or Solenoid on high-speed side of reducer

Thrustor or Solenoid on high-speed side of reducer and/or post or hand brake on tail shaft

Capacity Decelerating force minus friction force*

Incline load tension minus friction force*

Decelerating force plus incline load tension minus friction*

Anti-rollbackNeed None Essential None

Type - - -Ratchet, band brake, or one-way clutch located on high-speed side of reducer and/or on head shaft

- - -

Capacity - - - Incline load tension minus friction* - - -


C. Restraint of Decline Conveyors

Brakes and anti-rollback devices apply to stopping and holding conveyors. Brakes are not used to continually absorb the power generated by a decline conveyor. They do not have the capacity to do so.

Decline conveyors in normal running are restrained by the motor, which serves as a generator when the belt and its load begin to force the motor above its synchronous speed. No special motor or special control of the motor is required to produce this result. It is entirely a matter of the speed at which the motor is forced to run. The more torque the overrunning belt exerts on the motor, the farther above its synchronous speed the motor is driven and the greater its output as a generator. It fails to restrain the belt only when it is forced to such a speed that its current output becomes too heavy and its over-load protection interrupts the circuit. In that case, the interruption permits the brake to set and stop the conveyor. This does not happen if the motor is properly selected at the start. Also, downhill conveyors have centrifugal switches that open the control circuit at predetermined overspeed and thus set the brake.

The output of the conveyor through the motor goes to the powerline. It is presumed that this line is sufficiently loaded at other points to absorb the output of the conveyor motor, which acts as a generator.

Water brakes have been used to absorb the power generated by downhill conveyors. This requires a cooling system f'or the water, and the power output of the conveyor is wasted as heat. Since a motor is required to operate the empty belt, it ordinarily would be better to make the motor large enough to absorb the loaded output of the conveyor. An exception might exist where a downhill belt was driven by an internal combustion engine. In this case, a water brake could be used to take over when the belt became regenerative.



Handbook of Conveyor & Elevator Belting - Goodyear P1

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