In connection with U.S. Bureau of Mines funded-study relating to continuous haulage from deep open-pit mines, the authors had occasion to study extensively, the use of high-angle conveyors as a means of hauling mine products continuously from the pit while maintaining optimum mine slopes to minimize total excavation. The many possible high-angle conveying methods investigated included skip hoists, bucket ladders, pocket belts, fin belts, cleated belts, sandwich belts, pipe belts, screw conveyors, slurry pipe lines, and others. The sandwich belt high-angle conveyor appeared to be the most operationally appropriate and economical solution for the mining industry. This paper presents the theory, historical development, state-of-the-art and evolutionary developments in sandwich belt high-angle conveyors.
HISTORY AND THEORY OF SANDWICH BELT CONVEYORS
LIMITATIONS OF CONVENTIONAL CONVEYORS
Conventional belt conveyors offer an economical method for transporting bulk materials at recommended inclination angles ranging from a low of 7 degrees for soda ash briquettes, to a high of 30 degrees for cinder concrete and ground phosphate fertilizer.
Typical recommended inclination angles for open-pit mine products such as excavated earth and blasted, primary crushed rock vary from 15 to 22 degrees with respective angles of repose from 29 to 44 degrees.
The conventional conveyor is often the most economical, reliable, and safe means of transporting bulk material. There are, however, many cases which strongly warrant an increase in the conveying angle. The West Germans used high-angle sandwich belt conveyors on some bucket wheel excavators to increase the depth of cut without increasing the boom length. A substantial cost savings resulted from reduced structural and mechanical components associated with the shorter boom length required to accommodate a high-angle boom conveyor. In open-pit mines with steep mine face angles, or in any case where the surface incline angle in the path of the conveyor substantially exceeds that of the recommended conventional conveyor angle, much excavation or elevated support structure is needed in order to accommodate the conventional conveyor.
In order to logically develop the theory for high-angle sandwich belt conveyors, it is important to understand the nature of the conveying angle limitation for conventional conveyors.
In a static case, a cohesion-less material on a rubber belt will begin to slide back when the incline angle of the belt surface just exceeds the angle of internal friction of the material or the friction angle at the material to belt surface interface, whichever is smaller. The angle of internal friction is equal to the angle of repose for such materials. Both the angle of repose and the friction angle for bulk materials on rubber will vary from one material to the next and will be affected, even for the same type of material, by the maximum lump size, the lump-size distribution and orientation on the conveyor cross section, and the shape that the particles or lumps take as a result of the reduction process, that is, blasting and the varying degrees and methods of crushing. Although there is much published information on angles of repose for various bulk materials, very little is available on the friction angles for various bulk materials on a rubber surface. H. Colijno lists coefficients of friction of material on rubber for six different fine-grain materials (p. = tan cb, where p is the coefficient of friction, and cl) is the friction angle). They vary from 0.45 for Ottawa sand with 2 pct moisture to 0.624 for bituminous coal fines with IS pct moisture. The respective sliding friction angles are 24.2 and 32.0 degrees. The corresponding angles of internal friction are not listed. An investigation by the U.S. Bureau of Mines” revealed much higher friction angles when coarse material was laterally displaced over rubber belting material in a “Large Direct Shear Tester.” In one set of tests, river run gravel of size 0.25 by 0.185, 0.375 by 0.25 and 0.5 by 0.375 in. resulted in friction angles at the material to belt interface of 38.5, 42.0, and 38.0 degrees respectively. The corresponding angles of internal friction are 43.5, 48.4, and 52.5 degrees. A second set of identical tests on 1 1/2 by 0 in. Upper Freeport coal, 1 1/2 by 0 in. Middle Kittanning coal, 3/4 by 1/2 in. dolomite, and 3/4 by 1/2 in. limestone yielded respective friction angles, between the material and rubber belt, of 34.8, 38.1, 40.6 and 33.7 degrees. The corresponding internal friction angles are not listed.
The recommended conveying angles are, in general, far below either of the friction angles mentioned. This is due to the dynamics induced in a moving belt conveyor, which result in relative motion between adjacent particles or lumps of the bulk material and between the material and carrying surface of the conveyor belt. The three major sources of dynamic effects are:
- The agitation of the material on the conveyor belt as it approaches and is carried over each successive carrying idler. This effect is amplified when belt tension is low and idler spacing is high, thus resulting in large belt sag between idlers.
- The acceleration of the material at loading or transfer. This results in relative slip and turbulence because velocity and direction cannot change instantaneously.
- The vertical impact at transfer points. Such impact is absorbed by the resilience of the belt and impact idlers and results in bouncing of the material. This adds to the turbulence at the loading area.
These effects are increased when the belt speed is high and even more so when the material handled is loose and contains large rounded lumps.
These effects are increased when the belt speed is high and even more so when the material handled is loose and contains large rounded lumps.
In 1934, A. Vierline performed tests on a belt conveyor of V-shaped or wedge-shaped cross section and showed that he could increase the conveying angle simply by increasing the troughing angle. The conveyed materials were brown coal, household briquettes, briquette fragments, and overburden. It was theorized that it did not matter whether the material to be transported is carried in a single wedge-shaped trough, using two-roll idlers, or is carried on a belt with parallel wedge-shaped grooves, as long as the size of these grooves is in a prescribed ratio to the size of the pieces of transported material.
His tests showed that the conveying angle could be substantially increased only when the troughing angle exceeded 45 degrees. Tests conducted with a grooved belt conveying overburden with a natural angle of repose of 40 degrees showed that a conveying angle of 38 to 40 degrees could be achieved. It is not known at what belt speed, belt tension, idler spacing, and material-size distribution these tests were performed. It was announced however that the grooved belt could convey material with mixed grain-size composition at conveying angles 4 to 5 degrees less than the natural angle of repose.
Extrapolation of these results would lead to the conclusion that a grooved belt is effective in suppressing the dynamic effects due to belt travel over the idlers and in possibly creating an added normal pressure, due to wedging action, at the material to belt surface interface. The conveying angle cannot, however, exceed the angle of internal friction of the material at the free surface.
SANDWICH BELT CONVEYORS
Conveying angles greater than the angle of internal friction can be achieved by a cover which, when pressed against the material, will create a hugging action to prevent sliding of the contact surfaces.
For a cohesion-less material one can idealize the situation as shown in Figure 1. The material is idealized as parallel layers spaced infinitely close.
For the case where the cover surface is free to follow the material as it slides back, sliding will occur when the tangential component of the material weight exceeds the frictional forces which resist it or
Wm sin a > (N + Wm cos α)μ’……….(1)
where:
μ = μm or μb whichever is smaller.
Wm, α, N, μm, μb, are as defined in Figure 1.
To achieve an inclination angle α, a normal lineal hugging load, N, must be exerted by the top surface such that:
N ≥ Wm ( sin α – cos α) ………. (2)
If the cover surface is fixed and it resists the motion of the material at the interface, then material will begin to slide back when:
Wm sin α >N (μ” + μ‘) + (Wm cos α) μ‘………. (3)
where:
μ” = μm or μ, whichever is smaller. μ, is as defined in Figure 1.
To achieve an inclination angle α;
N ≥ Wm μ‘ ( sin α- cos α) ………. (4)
If μ’ = μ” , substitution into the above equation shows that the normal hugging load, N, needed to prevent backsliding, is only half of that required in the previous case (see Equation 2).
When the above principles are related to sandwich belt conveyors, the bottom surface represents the carrying belt, while the top surface represents any cover surface such as a pressed belt, a weighted belt, a link chain matt, etc. For a covering surface which depends on its self-weight, the normal lineal hugging load N becomes the normal component of the lineal weight of the cover surface. The second set of equations, 3 and 4, clearly show that the required hugging load is much less if both surfaces are driven at the same speed.
In his 1958 review of patented high-angle conveying methods, P. Rasper breaks down the sandwich belt approach into three different categories:
(a) Cover belt acting by its own weight.
(b) Cover belt with additional pressure.
(c) Cover belt with carriers.
Figure 2 illustrates a category (a) solution which was patented in West Germany in 1953. It consists of a carrying belt conveyor on three-roll, 30-degree troughing idlers with special spring-suspended impact idlers at the loading area, a cover surface which consists of a link chain matt, and a cleated belt to drive the cover surface. Special grooved return pulleys are used to carry the link chain cover matt.
During operation the material loads onto the bottom horizontal portion of the carrying belt, and is carried into the sandwich where the chain matt cover hugs the material by virtue of its self-weight, and prevents it from rolling or sliding back. The material is then elevated in the sandwich and is discharged at the top where the two surfaces separate.
The chain matt could conform easily to the irregularities in the material profile and this was thought to be an attractive feature. It was noted that a cover surface which depends on its self-weight can be uneconomical at higher conveying angles because the required lineal weight of the covering surface increases drastically. Some elevating conveyors were built according to these ideas.
Figure 3 illustrates a category (c) solution which was also patented in West Germany in 1953. It consists of a carrying belt conveyor on three-roll troughing idlers and a cover belt with flexible carriers consisting of flexible chain or rope secured to the belt at prescribed intervals along the length. It requires a loading belt in order to place the material into the sandwich.
During operation, the material is placed into the sandwich by the feed conveyor and it is immediately covered by the carrier. The cover belt and carriers hug the material, by virtue of their self-weight, while it is elevated to the discharge point at the top. The main selling feature of this method was the claim that the cover surface was self-cleaning because of the flexibility of the carriers. No elevating conveyors were ever built according to this design.
In his book The Bucket Wheel Excavator, L. Rasper discusses the elevating conveyor illustrated in Figure 4. It is a category (b) solution which was patented in West Germany in 1954. It consists of a carrying belt conveyor on three-roll troughing idlers and a cover belt which is pressed onto the material by rubber tires distributed across the belt and spaced at large intervals along the conveyor length. A feeder belt is used to load the conveyor.
During operation, the material is fed onto a low-angle, uncovered portion of the carrying belt at the bottom of the elevating conveyor, and is carried into the belt sandwich. The cover belt, which is pressed by the rubber tires, hugs the material as it is elevated to the discharge point at the top. According to L. Rasper, this solution was used on several occasions on bucket wheel excavators of the 1950s.
As a result of his survey, P. Rasper lists three important operational requirements of high-angle conveyors. The first pertains only to their use in bucket wheel excavators, while the second and third are important in any application.
- The high-angle conveyor must provide for easy removal or lifting of the cover belt when the conveyor is to be operated at shallow angles.
- The system must lend itself to easy and quick repair of the
- Any high-angle conveyor must lend itself to easy cleaning.
Requirement 1 illustrates a recognition of additional wear on the belt and other components when operating with the extra pressure from the cover surface. On a bucket wheel excavator boom, such a conveyor will operate at high angles only when it makes the extreme high and low cuts.
During operation, a bucket wheel excavator can pick up large boulders of different sizes and shapes, and tramp iron such as beams, plates, etc., from previous underground mines. These can cause tears in the belt.
Requirement 2 recognizes that it is impossible to guard against this so the effort must be to minimize the consequential downtime. For a regulated and more predictable material and size distribution, that is, the discharged material after primary crushing, good design can reduce the frequency of damage. The shape of the lumps cannot, however, be controlled. P. Rasper suggests that the repair be made by hot or cold vulcanizing and, on this basis, rules out the use of fins or cleats.
Requirement 3 is based on a realization that belts get dirty and must be continuously cleaned if they are to maintain their alignment. Buildup of material on the belt and other surfaces also leads to additional drag and premature wear of all support and drive components. He states that the most popular way to clean belts on bucket wheel excavators is by screw-type return idlers, and this requires the use of smooth surface belts. This requirement still holds true for the many belt cleaning methods available today.
On this basis he concludes that only solutions of categories (a) and (b) lend themselves to the operational requirements and thus warrant further study and development.
L. Rasper also points out that most of the development and use of high-angle conveyors on bucket wheel excavators occurred in the 1950s and that the bucket wheel excavators commissioned in the 1960s had abandoned them in favor of higher tonnage rates and reduced belt wear of conventional conveyors. This reduced the cutting depth of the machine, but operational experience showed that deep-cut cohesive soils and excavated lignite permitted conveying angles of up to 23 and 27 degrees, respectively, with conventional conveyors.
More recent developments in the category (a) solutions include the Retainer Belt by Stephens-Adamson and the Overlay Conveyor by R.A. Hansen Company. Both claim the capability to convey material at angles up to 45 degrees.
The Retainer Belt, as illustrated in Figure 5, consists of a conventional carrying conveyor on three-roll troughing idlers and a heavy cover belt with smooth rubber surface. The cover belt is shot with lead for the extra weight needed for the normal force component.
During operation, the material loads onto the uncovered conventional conveyor, at a nearly horizontal conveying angle, and enters the belt sandwich prior to a gradual increase in conveying angle. The sandwiched material is conveyed at the high angle to the discharge point where the belts separate.
The Retainer Belt conforms well to the second and third operational requirements listed by P. Rasper. Other than the special cover belt, it uses only conventional conveyor components. The high cost of the special cover belt, however, makes the Retainer Belt expensive at high angles. There is no apparent capacity limitation other than those which govern conventional conveyors.
The Overlay Conveyor is illustrated in Figure 6. It also conforms well to the second and third operational requirements set by P. Rasper, but it does not address the need for hugging pressure to keep the material from sliding back. It uses an ordinary (non weighted) cover belt which is driven along with the carrying belt by mechanical coupling. The carrying belt is supported on three-roll troughing idlers. A specially suspended single-roll idler presses the cover belt onto the carrying belt at the entrance to the belt sandwich. A speed-up belt is used to, feed the high-angle conveyor.
During operation, the speedup belt throws the material into the loading area and its momentum carries it into the sandwich entrance. From that point it is carried in the belt sandwich until it is discharged at the top where the belts separate.
It is argued that the ordinary non weighted cover belt is sufficient to keep the large lumps from rolling back and the rest of the material will not slide back. It is also argued that the speedup belt will throw the material into the loading area at a prescribed speed and angle of incidence so that the momentum of the material will carry it into the belt sandwich with a minimum amount of turbulence. This justifies loading an uncovered conveyor at the high angle.
Two test units were built to verify the theory. The first used a 20-in. cleated carrying belt with an 18-inch smooth cover belt and conveyed wet sand with 10 pct rocks at 45 degrees while the second used two 60-in. smooth belts and conveyed wet sand with 10 pct rocks at a conveying angle of 40 degrees. These test units were successful in conveying wet sand at high conveying rates and wet sand with 10 pct rocks at reduced rates. When the rocks exceeded 10 pct, they would roll and bounce in the loading area without entering the sandwich. The ability to convey a limited type of material, without additional weight on the cover belt, illustrates the advantage of driving both belts and indicates that the effective friction angles within the sandwich may be quite high.
In light of the state-of-the-art and current developments in sandwich belt-type high-angle conveyors and the requirements of the present development, two new categories (d) and (e) are added to the possible solution types as listed by P. Rasper.
The current solution categories are then as follows:
(a) Cover belt acting by its own weight.
(b) Cover belt with additional pressure.
(c) Cover belt with carriers.
(d) Belt sandwich with prying resistance by virtue of the transverse stiffness of the two belts.
(e) Belt sandwich with radial pressure by virtue of the belt tension and the conveyor profile geometry.
The Loop Belt by Stephens-Adamson represents a category (e) solution. It is illustrated in Figure 7.
With this concept, a non weighted inner belt loop is pressed against closely-spaced troughing idlers by an outer belt loop. The outer belt is straight at the loading region and supported on troughing impact idlers. When the conveyed material enters the belt sandwich, the carrying strand (outer belt loop) is no longer supported by idlers. As the belt sandwich enters the vertical curve, the belt tension and the radius of curvature are such that sufficient radial pressure is produced to overcome the weight of the belt and the conveyed material. The material is thus conveyed in a sandwich made up of the suspended carrying belt and the idler-supported inner belt loop. This sandwich loops around, approximately 155 degrees, to the discharge point. The lineal hugging load produced by the outer belt is determined by equation 5. As the conveying angle increases, the vertical radius of curvature and the outer loop-belt tension must provide sufficient radial pressure to support and hug the suspended material in order to convey it up to a 90-degree angle. Above the 90-degree angle, the material begins to sit on the supported inner belt loop; and the outer belt loop becomes the cover belt which supplies the hugging force. The material is conveyed in this manner to the discharge point. Typically, the outer belt loop is driven while the inner belt loop is forced to follow by the frictional drag of the conveyed material. There has been occasion to drive both belts by a load-sharing drive arrangement which utilizes the combined tension rating of both belts in order to maximize elevating height.
Pr = T ……….(5)
where:
T = Outer belt tension at the point in question on the conveyor profile.
R = Radius of curvature corresponding to the belt tension.
Pr = The corresponding lineal load applied by the outer belt.
The Loop Belt has had great success in self-unloading ship applications and several outdoor installations. It has achieved 10,000 tons per hour, in conveying iron ore pellets and coal, and lifts up to 150 feet. Typical belt speeds are 860 to 1,200 feet per minute. A special low-modulus fabric belt is used to achieve tight vertical curves with a troughed belt.
The Beltavator, by Stephen-Adamson, is a hybrid which represents a solution of categories (d) and (e). It is illustrated in Figure 8. The Beltavator is an extension of the Loop Belt. It is identical to the Loop Belt up to a conveying angle of 90 degrees. At this point, it conveys the material vertically for a desired elevating height and it again becomes identical to the Loop Belt above the 90-degree conveying angle. It is the straight vertical portion of the conveyor profile that qualifies as a category (d) solution. On the straight vertical portion, the belt sandwich is held together by closely-spaced, staggered-edge rolls which press the belt edges to keep the sandwiched material sealed between. The hugging pressure required to convey the material at 90 degrees is provided by the prying resistance of the two belts as the material is introduced into the sandwich. The prying resistance is by virtue of the transverse stiffness of the two belts. The required transverse stiffness imposes a capacity limitation on the Beltavator. As the belt width requirements become large, in order to achieve higher tonnage rates, increasingly thicker multi-ply fabric belts are required to provide adequate transverse stiffness. Such considerations limit the practical belt width to a maximum of approximately 36 inches, and maximum capacity is approximately 1,000 tons per hour when conveying dense material. Only the outer or receiving belt is typically driven as was the case in the Loop Belt.
The Beltavator, too, has had great success within its capacity range. Its conveying profile can follow a “C” Belt Path,” a “Z Belt Path,” or an “L Belt Path,” as shown in Figure 8.
It is important to point out that solutions which fall into categories (d) and (e) represent a significant advance in the historical development of the sandwich-type high-angle conveyors. All of categories (a) through (e) address the need for hugging pressure which increases the frictional resistance to relative movement between adjacent particles of the conveyed material and between material and belt surfaces. The significance is that solution categories (d) and (e) do not require added weight or a pressing mechanism.
In the category (d) solution, the hugging pressure results from prying resistance and the transverse stiffness of the two belts, this transverse stiffness is utilized to an advantage, but it exists whether or not it is utilized.
In the category (e) solutions, the hugging pressure results from radial force by virtue of the tangential belt tension and the curved geometry of the conveyor profile. Here the belt tension is used to an advantage by the selection of a particular conveyor profile geometry. The belt tension at any point on the profile is related to the tension at the tail pulley, the lineal material load, and the height above the tail pulley. As with the belt transverse stiffness, it exists whether or not it is used to an advantage by the selection of the appropriate conveyor profile geometry.
Additional implications in category (e) solutions are discussed later when addressing the vertical radius of curvature requirements as specified by CEMA.
It may seem surprising at first that Stephens-Adamson typically chooses to drive only the carrying belt in the Retainer Belt, the Loop Belt, and the Beltavator when it is clear from equations 2 and 4 that the required hugging pressure is greatly reduced by exerting drag at both the carrying and cover surfaces. A closer and more realistic look at the belt sandwich model will reveal that both surfaces do indeed exert drag on the material, even though only one belt is driven.
The Belt Sandwich model, illustrated in Figure 1, is very instructive but it is not totally accurate. It assumes that the cover surface contacts only the material, but the edges do not touch the carrying belt. It is known that lateral movement of the cover surface during operation will cause the edges to bear, intermittently, on the carrying belt; thus losing a portion of the hugging load directly to the carrying belt and support idlers, while uncovering the material at the other edge. More realistically, a minimum distance from the belt edge to the material is required in order to assure that the material is always covered and does not spill out. On this basis, Stephens-Adamson chooses to derate the capacity of the sandwich type conveyors (as compared to CEMA capacity recommendations for conventional conveyors). This assures large edge distances and, thus, a sealed envelope.
A new, more realistic model (Fig. 9), illustrates the interplay of forces. The minimum normal hugging load, Nm; which must be exerted on the material in order to prevent backsliding, if both surfaces resist motion, is expressed by equation 6.
(min) Nm = Wm sin α cos a) ………. (6)
This follows from equation 4. The drag which must be exerted on the material by the top surface must be counteracted by the frictional drag between the top and bottom surfaces at the edges, as expressed by equation 7.
(min) N, µr = (min) Nm μ, ………. (7)
The minimum required total normal load, N, can be expressed by combining equations 6 and 7 to obtain equations 8 and 9.
(min N = (min) Nc + (min) Nm = μ” + 1) (min) Nm ………. (8)
(Min) N = (μ” + 1) Wm μ” sin α – cos α)……….(9)
If both carrying and cover surfaces are of rubber conveyor belting, then μ‘ = μ“. If = μ’ = μ”, equation 9 reduces to equation 2. It will be shown later that this is, in fact, a reasonable assumption. Equation 2 assumes that only one surface is driven.
At first impression it would seem that nothing has been gained buy using the new more realistic model, because the required total normal load, N, is the same when only one belt is driven regardless of which model is used. What is important is the recognition of a more realistic behavior and that only the portion of the normal load which bears directly on the conveyed material actually contributes to an increase in the impact on each successive idler as the material is carried along with the belt, and thus, only this portion contributes to any increase in belt wear. The required edge distance corresponding to the assumption that μe = μ‘ = μ” is b/4, where b is the belt width.
What then is the advantage of driving both surfaces? If the type of material to be handled and the environmental constraints do not require a large edge distance, then the total required additional normal load, N, can be reduced. If there are circumstances which would cause doubt that the required counteracting edge drag could be developed, that is, μe < μ”, then it would be advantageous to drive both surfaces. In order to maximize total elevating height, it is advantageous to drive both surfaces by a load-sharing drive arrangement so that the combined tensile strength of both belts is utilized.
In light of the more current solutions in sandwich belt conveyors, a new set of operational requirements is established. This set retains requirements 2 and 3 as originally listed by P. Rasper (now numbered 4 and 5), but omits his requirement 1 because it is applicable only to high-angle conveyors when used on bucket wheel excavators.
The combined operational requirements are:
- The receiving end of the steep-angle conveyor must insure that the turbulent material at the load point is settled quickly and experience no backsliding prior to entering the sandwich.
- The hugging pressure exerted on the conveyed material and, thus, the additional pressure on the cover surface must be applied by a “soft” loading system which minimizes load concentrations.
- The cover surface must be a floating surface which will not obstruct the flow of lumps larger than anticipated or oriented unfavourably.
- The system must lend itself to easy and quick repair of the
- Any high-angle conveying system must lend itself to easy cleaning.
Requirements 4 and 5, which have already been discussed, preclude the use of carriers, and makes it impossible to cover the material immediately as it is fed into the high-angle conveyor. The Overlay Conveyor by R.A. Hansen Company(11) was loaded at a 40-degree angle by using a speedup belt to throw the material into the loading area and allowing the momentum to carry it into the sandwich. This was apparently successful in conveying wet sand, but many difficulties arose when attempting to convey wet sand with 10 pct rocks. It was found that the orientation of the speedup belt, with respect to the loading area, was very critical and the optimum orientation varied with variations in the mix of the bulk material. Rocks alone could not be transferred in this manner because they would bounce and roll back. But even more basically, such a system, which depends on the momentum and angle of incidence of the fed material, could not tolerate an emergency stop of the conveyor. It would be impossible to re-accelerate any material remaining at the loading area after the shutdown. Requirement 1 recognized the problems associated with loading a sandwich belt conveyor at a high angle. On this basis, it is recommended that the conveying angle onto which material is loaded should not exceed 10 degrees, and the material should enter the sandwich at a maximum conveying angle of 18 degrees.
Compliance with requirement 2 will minimize the additional wear on the belts and other components due to the additional hugging load. This is very important because L. Rasper cites the reduced belt wear of conventional conveyors as one reason why the bucket wheel excavators of the sixties had abandoned the use of high-angle boom conveyors.
The design must convey material which is discharged from a primary crusher. The crusher setting only assures that one dimension will not exceed the prescribed value. If the feed into the crusher consists of slabby material, then a minus 8-in. setting could result in slabs of dimensions of up to 8 by 12 by 24 inches. The loading area should be designed to assure that such slabs are oriented favorably. Requirement 3 assures that over sized or unfavorably oriented material will not encounter pinch points and will result only in local lifting of the cover and possible spillage of the material. It should be emphasized that over sized lumps and slabs of the dimensions mentioned are not to be considered part of the normal operation. If slabby material is to be crushed, it may become necessary to reduce the crusher setting.
CONSTRAINTS ON VERTICAL RADIUS OF CURVATURE
Operational requirement 1 carries with it severe implications. After the material is loaded, the conveying angle must be increased from 10 degrees to a much higher angle. In open-pit mine application, the ultimate conveying angle is dictated by the slope of the mine face. Such an angle increase cannot be instantaneous with a troughed belt. The angle change must be sufficiently gradual so that no part of the troughed belt is subject to buckling or over stress. Within these constraints we must strive to increase the conveying angle within the shortest possible distance. Figures 10, 11, and 12 illustrate why this is so.
If α represents the highest mine face angle which is attainable because of slope stability or other considerations, then a multi-run high-angle conveyor system may be incorporated into the mine, as illustrated in Figures 10 and 11.
Figure 10 illustrates the adaptation of the conveyor system without additional excavation. The consequences of a long transition radius of curvature are two-fold. First, the ends of each module divert farther from the mine face, thus increasing the structural support requirements. Secondly, the ultimate conveying angle α”, corresponding to the module with long transition radius, is greater than α’, which corresponds to a module of short transition radius. Thus, a higher ultimate hugging force is required to prevent the conveyed material from backsliding.
If additional local excavation is favored in order to reduce the structural support requirements and to keep the conveyor profile within easy access from the mine benches, then the conveyor system is adapted to the mine as shown in Figure 11. A long transition radius of curvature results in increased excavation requirements.
If a single-run conveyor is used to carry the material out of the mine, as in Figure 12, then a long transition radius of curvature requires that the conveyor loading area extends further into the mine. In many cases it is more efficient to locate the loading area several bench depths above pit bottom. In such a case, the imposition of the conveyor loading area makes it difficult to recover the material below. Depending on the slope stability requirements, the amount of irrecoverable material can be substantial.
The equations to determine the allowable vertical radius of curvature are listed by CEMA and derived here. The final form of the equations as derived are equivalent to the CEMA equations but are presented in a more instructive form.
The vertical belt curve must be designed so that it will not cause buckling nor over stress at any part of the belt cross section. CEMA, in fact, requires that the minimum allowed tensions anywhere on the belt is 30 pounds per inch of width.
When any section of linearly elastic material is subject to simultaneous tension, Tc, and bending, M, which is induced by the curvature 1/r as shown in Figure 13, the stresses due to the independent forces may be superimposed. The model assumes that the curve is smooth and that plane sections remain plane. When applied to a conveyor belt section, this model predicts the stresses accurately at points sufficiently far from the ends of the curve, if the trough depth and width are much less than the length of the transition curve and the support idlers are very closely spaced when compared to the radius of curvature. Having made these assumptions, the bending moment can be related to the radius of curvature by the following equation:
M = El……….(10)
where:
E = modulus of elasticity (lbs/in.)
l – sectional moment of inertia (in.)
Superposition of stresses may then be expressed as follows:
f = fa + fm = Tc + E 1……….(11)
where:
f, fa, fm, Tc, A, r S are as defined in Figure 13
Recognizing that I/S = y and substituting into equation 11 yields:
f = fa + fm Tc + E y……….(12)
where y is the distance from the neutral axis to the extreme outer fibers (see Fig. 13).
This equation can be applied to a multi-ply conveyor belt of arbitrary geometry by making the following substitutions:
— Replace the cross-sectional area A (in.2) in equation 12 by the product of the belt width and the number of piles. (b) by (p) (in.-ply)
— Replace the elastic modulus E (lb/ in.2) by the commercially-listed belt modulus Bm (lbs/in.-ply).
Equation 13 then follows from equation 12.
f = fa + fm = Tc + Bm y (lbs/in.-ply) ……….(13)
The units of stress f, fa, fm are now expressed in lbs/in.-ply.
To prevent over stress of the bottom fibers, the combined stresses must not exceed the working tension rating, Tr, of the belt. This means that the following equation must be satisfied.
Tr ≥ Tc + Bm y’ ………. (14)
where:
Tr = Working tension rating of the entire belt (lbs)
y’ = distance from the neutral axis N.A. to the extreme bottom fibers (in.)
solving for r yields the following:
r ≥ b Bm p y’………. (15)
To maintain a minimum tension of 30 lbs/in. on the top fibers, the following equation must be satisfied:
30 ≤ Tc – Bm y” ………. (16)
where:
y” = distance from the neutral axis N.A. to the extreme top fibers (in.)
Solving for r yields:
r ≥ b Bm p y”………. (17)
Next, consider a belt which is carried on three-equal-roll idlers of troughing angle. It is assumed that the belt is divided into three equal parts, as illustrated in Figure 14.
If the vertical curve under consideration is concave (that is, belt edges are at inner side of curve), then y’ = (b/9) sin Φ and y” (2b/9) sin. Equations 18 and 19 follow when these identities are substituted into equations 15 and 17, respectively, and the equations are divided by 12 so that R is the radius of curvature in feet rather than inches.
R ≥ b2 Bm p sin (to prevent over stress of middle when curve is 108 (Tr-Tc) concave)………. (18)
R ≥ b2 Bm p sin (to prevent edge buckling when curve is 54 (Tc-30b) concave)………. (19)
If a convex vertical curve is considered (that is, belt edges are at the outer side of the curve), then y’ = (2b/9) sin 0, and y” = (b/9) sin 0. Equations 20 and 21 follow when these identities are substituted into equations 15 and 17, respectively, and the equations are divided by 12 so that R is in feet rather than inches.
R ≥ b2 Bm p sin co (to prevent edge over stress when curve is 54 (Tr-Tc) convex) ………. (20)
R ≥ b2 Bm p sin e (to prevent buckling of center when curve is 108 (Tc-30b) convex) ………. (21)
Equations 18, 19, 20 and 21 are equivalent to the CEMA equations and apply to multi-ply fabric belts. For steel cord belts, equations 18, 20, and 21 are applicable if p = 1 and the elastic modulus Bm is in (lbs/in). When the curve is concave, the belt edges of a steel cord belt are allowed to buckle to the extent that it is not detrimental to the operation. To reflect this, the right-hand side of equation 19 is divided by 2.5, thus yielding equation 22.
R ≥ b2 Bm p sin Q (allow edge buckling in steel cord belt of 54 (Tc-30b) 2.5 concave curve)………. (22) where p = 1 and Bm is in (lbs/in)
Equations 18 through 22 reveal the governing parameters as they relate to the allowable radius of curvature. The belt width, b, and the troughing angle, 0, must be established ahead of time so that they are compatible with the capacity requirements and maximum lump size of the material to be handled. For any solution of category (e), it is established that the trough depth shall not be less than the maximum lump size (that is, for belt on three-equal-roll idlers, then (b/3) sin (P> maximum lump size). With belt width and troughing angles established, the remaining variables are the composite elastic modulus of the belt, Bm p (lb/in.), the rated working tension, Tr (lbs), and the operating belt tension at the point under consideration, Tc (lbs). For a concave curve, equations 18 and 19, applicable to multi-ply fabric belts, or 18 and 22, applicable to steel cord belts, must be satisfied simultaneously. Tc may be increased by an increase in tail tension in order to counteract the buckling tendency but not to the point where the working tension rating, Tr, is exceeded. The same arguments apply to equations 20 and 21 for convex curves. In order to minimize the radius of curvature without violating the governing equations we must seek a belt which maximizes the ratio of the rated tension to the elastic modulus (maximize (Tr / Bm p)). However, very low values of Tr cannot be acceptable because this would severely limit the elevating height of any single run. The nylon-by-nylon reinforced fabric belts seem to offer the best solution of the commercially-available multi-ply fabric belts. In addition, it is possible to order a special nylon-by-nylon belt of reduced elastic modulus, Bm. The cost of such a special belt is higher, but installations in limited space may require a very short vertical radius of curvature and, thus, warrant the extra cost. The Loop Belts installed by Stephens-Adamson on self-unloading bulk carrier lake ships incorporate a special low-modulus belt. An added benefit is that the tail tension on the carrying belt, required to suspend and hug the material at the sandwich entrance of the Loop Belt, is reduced by the reduction of the radius of curvature as dictated by equation 5.
If we consider a standard nylon-by-nylon fabric belt, 60 inches wide, carried on equal-roll troughing idlers at 30-degree troughing angle and subject to a convex transition curve, we could expect a minimum allowed radius of curvature in the range between 50 and 80 feet. The variation is due to the different standard elastic moduli listed by the various manufacturers. One could expect radii of curvature equal to one half of this value if a special low-modulus belt is used. With a steel cord belt, the minimum allowed vertical radius of curvature would typically exceed 10 times that of the standard nylon-by-nylon fabric belt. In open-pit mine applications, the steel cord belt is feasible only in a single-run application with the substantial penalty of making it difficult to recover the material below the transition curve, as illustrated in Figure 12.
When large transition curves, such as required for a troughed steel cord belt, are used to increase the conveying angle from 10 degrees at the loading area, to the ultimate angle of the mine face; a category (e) solution is not possible along the transition profile (see Fig. 16) because the corresponding belt tensions, as dictated by equation 5, are not practical. If category (a) or (b) solutions are employed, as shown in Figure 15, then an additional constraint applies to the transition curve. Because the carrying belt is supported on troughing idlers and is uncovered from the 10-degree angles at the loading area to the 18-degrees angle at the sandwich entrance, uplift of the belt in this region must not occur when starting or running empty. Typical operating tensions in this region will probably not cause uplift of the empty belt but,. if it is necessary, the added load of the covering system may be applied at angles lower than 18 degrees or prior to the start of the transition curve in order to insure that uplift does not occur.
Transition profiles of short radii of curvature, such as may be obtained with nylon-by-nylon multi-ply fabric belts, lend themselves to a category (e) solution. This is illustrated in Figure 16. The belt tension required at the tail end to suspend the material and carrying belt, as dictated by equation 5 is low enough to make this solution very attractive.
Utilizing 2 nylon-by-nylon fabric multi-ply belts, driven by a load-sharing drive system, net elevating heights of up to 350 feet could be obtained without over sizing the belts specifically to maximize lift. A single-run conveyor, utilizing two steel cord belts driven by a load-sharing drive system, could achieve practical elevating heights of up to 900 feet. In mines with depths exceeding 900 feet, the high-angle conveyor system could consist of fabric belt conveyor modules of short transition radius, arranged as shown in Figure 10 or Figure 11; or a single-run approach, as in Figure 12, utilizing steel cord belts, or a combination of both. The single-run approach is attractive because it eliminates the need for additional intermediate transfer points and reduces the number of pulleys and drive components. For a multi-level operation with conveyors of varying lift required, any single-run lift exceeding 350 feet would require steel cord belts, and thus a long transition profile, causing an imposition into the mine pit which would limit material recovery.
A strong desire to achieve lifts greatly exceeding 350 feet, without the penalty of a long transition radius, led to the investigation of a method using a steel cord carrying belt and a fabric cover belt with spring-mounted pivoting wing rolls at the idlers of the transition curve. The operational characteristics of this method proved unsatisfactory and it was not pursued further.
EVOLUTIONARY CONCEPTS FOR FUTURE DEVELOPMENT
Before introducing new concepts in sandwich belt high-angle conveyors, it is appropriate to establish criteria which recognize the positive and negative features of past developments, and serve as the basis for evolution of these into operationally-superior new concepts.
CRITERIA FOR DEVELOPMENT OF SANDWICH BELT CONVEYORS
All new developments in sandwich belt high-angle conveyors must address the theory and constraints discussed and should judiciously select and incorporate the desirable features of past developments while omitting, where possible, those features which are not considered desirable.
In addition to the five operational requirements already listed, new developments in sandwich belt high-angle conveyors should conform to the following specific design criteria. Criteria 4, 5, and 6 apply specifically to hard-rock material discharged from a primary crusher, set at 8-in. maximum lump size.
- Design coefficients of friction at the belt surface to material and belt surface to belt surface interfaces are taken as .5 which corresponds to a friction angle of 26.6 degrees.
- The belt width is sized so that the cross-sectional area of material, corresponding to a uniformly loaded conveyor operating at capacity, is less than the cross-sectional area recommended by CEMAII) for a conventional conveyor with 0-degree surcharge angle.
- The trough depth of the idler-supported belt in a category (e) solution is equal to or exceeds the maximum lump size.
- Belt speed is approximately 700 feet per minute and does not exceed 750 feet per minute.
- All carrying idlers are of three-equal-roll troughing type and all idlers conform to requirements of CEMA Class E7.
- Minimum wear cover thicknesses are:
- 3/8 inch carrying belt
- 1/4 inch cover belt
- Grade of Rubber: RMA 1 (Rubber Manufacturers Association)
The design coefficient of .5 between conveyed material and rubber surface, as established in Item 1, reflects the authors’ best estimate in light of the limited amount of data available in this area.
Because the material is totally bounded in the sandwich, local impact as the material is carried into and over successive carrying idlers is, to some extent, cancelled by the resisting reaction of the cover belt. The presence of the hugging force and elimination of the rolling tendency of the larger lumps, which rise to the top surface, assure a nearly static frictional resistance to movement at the material to belt surface interface. This value is thought to be conservative; but because of a lack of data in this regard, it provides a reasonable design criterion.
A design friction coefficient of .5 for the interface between two belt surfaces, as established in Item 1, is consistent with the values used in calculating frictional development in belt-on-belt-type intermediate drives. It takes into account surface wear and the introduction of dirt and moisture.
It is very convenient to have equal values for these friction coefficients, because the additional pressure to transmit relative drag between the two belts, in addition to that required to hug the material, does not depend on the material-free edge distance. This means the prediction of an exact configuration of the sandwich cross section during operation is not necessary. This may be verified by considering equation 9 and reviewing the discussion which precedes and follows it. This becomes important when considering a load-sharing drive arrangement.
Item 2 is an artificial way of achieving the desired cross-sectional configuration for a belt sandwich operating at the rated capacity. A scaled representation is shown in Figure 17. It can be seen that such a configuration yields ample edge distance to assure a tight, leak-proof seal and a margin for overload. It should be noted that Item 2 represents a de-rating of the belt cross section by approximately 50 pct when compared to a conventional conveyor with a surcharge angle of 25 degrees (as recommended by CEMA for conveying hard rock, lumpy material).
Item 3 reflects the experience of Stephens-Adamson with the operation of Loop Belts. It assures a relatively-smooth hugging pressure distribution without excessive local stretching anywhere on the belt. As noted, this criterion applies specifically to those methods which incorporate category (e) solutions. It turns out that all methods which use a short transition radius to increase the conveying angle from the loading area to the mine face angle do, indeed employ a category (e) solution.
Item 4 sets a conveyor belt speed which will assure smooth operation and extended life of the components. It reflects CEMA recommendations and the experience of industry, domestic and foreign, in conveying hard lumpy material.
Item 5 establishes that the most rugged idlers (the highest CEMA Class) shall be used in accordance with the requirements of continuously conveying hard and lumpy material in a harsh outdoor environment.
The RMA grade and minimum rubber cover thickness listed in Item 6 are as recommended by belt suppliers.
THE MECHANICALLY-PRESSED SANWICH CONVEYOR
The Mechanically-Pressed sandwich conveyor is illustrated in Figures 18 and 19. As noted in previous discussion, conveyors which relied on a self-weighted cover belt did not prove to be economical for conveying angles approaching 45 degrees. The category (b) Mechanically-Pressed conveyor, illustrated in Figure 4, had limited success on bucket wheel excavators but was eventually abandoned, for reasons which included the accelerated wear of the belt and other components. This is not surprising because the total required hugging load is applied as large discrete loads spaced at large intervals along the conveyor. Operational requirement 2 addresses this problem by calling for a soft loading system which minimizes load concentrations.
The Mechanically-Pressed sandwich conveyor fulfills all of the established operational requirements. It is as well suited to a modular approach, utilizing nylon-by-nylon fabric belts of short vertical radius of curvature, as illustrated in Figure 18, as it is to a single-run approach utilizing steel cord belts, as illustrated in Figure 20.
MODULAR APPROACH
The modular approach combines a category (e) solution at the transition curve with a category (b) solution on the straight run. The hugging load on the straight-run portion is by totally equalized standard impact idler rolls. The equalization is to achieve a soft loading system, as established in operational requirement 2. The pressing force is by deflection of the loading springs and therefore increases as required with increasing conveying rate.
Design trials at 6,000 tons per hour for simultaneous requirements of conveying at mine slope angles from 40 to 50 degrees achieved an elevating height of 330 feet. Two 84-inch standard multi-ply nylon-by-nylon fabric belts are supported on 20-degree troughing idlers along the transition curve of 60 feet radius of curvature, and carried on 35-degree troughing idlers on the straight portion. A load-sharing drive arrangement is used to maximize elevating height.
The conveyor profile lends itself to support by a straight truss with intermediate portals and access from any bench level. The simple geometry and proximity to the mine face facilitates fabrication and erection. Such a unit is self-contained because belt tensions are taken through axial compression of the main truss and no external belt anchorage is required.
SINGLE-RUN APPROACH
A single-run, Mechanically-Pressed sandwich conveyor was designed at 6,000 tons per hour for a 50-degree mine slope angle. Two 72-in. standard steel cord belts are carried on 35-degree troughing idlers and driven by a load-sharing drive arrangement to achieve an elevating height of 750 feet. This is illustrated in Figure 20. This is entirely a category (b) solution. Use of steel cord belts make it necessary to use a 1,000-foot transition radius of curvature and, thus, the imposition into the mine pit.
Because of the much increased conveyor length, large foundations are required at each end to decouple the high belt tensions from the remaining structural support. This might actually result in a support arrangement which is more economical in fabrication and first erection, but mobility is not as good.
THE PNEUMATICALLY-PRESSED SANDWICH CONVEYOR
Although the Mechanically-Pressed sandwich conveyor is a great improvement over the solution of Figure 4, it can be seen in Figure 19 that local wear patterns on the cross section would result at the edges of the pressing rolls. The consequences of such wear patterns, as they affect the belt life, need further investigation. The Pneumatically-Pressed sandwich conveyor of Figure 21 improves on this situation by applying the hugging load onto the cover belt by a soft air cushion.
Otherwise it is identical to the Mechanically-Pressed sandwich conveyor and lends itself equally well to a modular approach; which combines a category (c) solution at the transition curve with a category (b) solution on the straight run, as to a single-run approach, which is entirely a category (b) solution.
THE AIR-SLIDE SANDWICH CONVEYOR
The Air-Slide sandwich conveyor attempts to take the Pneumatically-Pressed sandwich conveyor one step further by supporting the carrying belt on an air plenum, rather than carrying idlers. A modular approach still uses loading impact idlers and support troughing idlers at the transition curve in a category (e) solution, but it uses a support air plenum and a hugging air plenum on the straight portion in a category (b) solution. This is illustrated in Figure 22. The single-run approach would use impact idlers at the loading area and a support air plenum and hugging air plenum elsewhere in a solution which is category (b) only.
Although the air pressure system offers the softest possible hugging load, there are many problems envisioned when it is used in a dirty outdoor environment.
Contaminated air can clog filters and wear down the high speed mechanical parts of the pressurization system. Failure of the pressurization system will result in a loss of the hugging load, and a backup system is required. Because the air plenum which presses the cover belt is in fixed position, it is necessary to minimize the air gap between the plenum skirts and the cover belt in order to keep the additional energy requirements, for the air system, within reason.
THE S-SHAPE SANDWICH CONVEYOR
The S-Shape sandwich conveyor is illustrated in Figures 23 and
24. This concept is truly evolutionary in that it follows logically from the Stephens-Adamson Loop Belt and Beltavator but avoids the shortcomings of each system. It is a category (e) solution which depends on the radial force generated by the belt tension and curvature of the profile. The constraints on the radius of curvature, as discussed previously, dictate the use of nylon fabric multi-ply belts.
As stated previously, the Beltavator is identical to the Loop Belt up to and above the straight vertical run portion. It is the straight vertical portion which limits the ability to achieve high conveying rates. Let’s consider the Z Belt Path Beltavator, as illustrated in Figure 8. If the straight vertical conveying portion is eliminated, and the remaining curved portions are joined, the essentials of the S-Shape sandwich conveyor remain. In addition, the relationship between belt tension and curvature is used optimally by continually varying the radius of curvature along the conveyor profile, as shown in Figure 23. The continual variation of the curvature is very important in achieving the maximum lift within the geometrical constraints, that is, conforming to the specified mine slope, and in achieving extended life of the conveyor belts and idlers by providing only the hugging force necessary at a given point on the profile.
The highest radial force requirement is at the bottom portion of the conveyor where the belt tension is lowest. From the last support idler of the receiving area to the sandwich entrance, the radial force must be sufficient to overcome the lineal weight of the conveyed material and the carrying belt with adequate provision for possible overload. As the conveying angle increases, the radial component of the material load decreases, but the tangential component increases, thus increasing the hugging pressure requirements to prevent backsliding of material. Calculations show that the same radius of curvature is consistent with the requirements from the loading area to the inflection or reversal point of the profile. Because the carrying belt tension is lowest in this area, equation 5 dictates that the radius of curvature is also shortest in this region. Above the inflection point, the carrying belt is supported by the idlers, and the cover belt provides the radial force to hug the material. The top belt tension at this point is chosen so that both belts will reach their working rating simultaneously at the drive pulleys where a load sharing drive arrangement is employed to take full advantage of the working tension ratings of both belts. As we proceed upward along the conveyor profile, the belt tension increases linearly with height; while the conveying angle gradually decreases, thus decreasing the hugging requirements on the conveyed material. This means that the radius of curvature increases rapidly, as shown in Figure 23. Failure to increase the radius of curvature, as shown, would result in excessive radial force which would result in accelerated wear of the belts and would require closer spacing of the idlers. In effect, the conveyor dictates its own geometry, and that which is consistent with the mine slope and the desired elevating height is determined by trial and error.
The ability to predict minimum operating belt tensions along the conveyor profile, which is essential in designing the profile geometry, is greatly dependent on the use of automatic take-up arrangements at the tail end of each conveyor belt and a load-sharing drive arrangement at the head end. The drive arrangement consists of a drive unit at each head pulley, with a control system which causes the drives to share the conveyor load proportionally to the relative belt stiffness, regardless of the conveying rate. The use of the automatic take-up arrangements allows prediction of the belt tensions with good accuracy at the tail end of the conveyor, while the load-sharing drive system allows prediction of belt tensions at the head end with an accuracy which is dependent on the characteristics of the load-sharing drive arrangement. A 10 pct tolerance in the relative belt tensions at the head end is considered acceptable and incorporated into the design of the components. With the belt tensions known at the head and tail ends of the conveyor, it must be assured that the increase in tension between the two ends, which is due to material load, self-weight of the belts, frictional drag, and acceleration-deceleration, is distributed to the two belts in proportion to their relative stiffness. To assure this, no relative movement must occur between the belt surfaces. The minimum hugging pressure must, therefore, develop sufficient drag between the top surface and the material and bottom belt edge surface so that slip will not occur due to the tension variation of the top belt, with respect to a stationary material and bottom belt surface. The minimum hugging pressure must also assure that the portion which is transmitted directly through the material is sufficient to prevent backsliding of the conveyed material. In the design, the ratio of the top belt stiffness to the combined belt stiffness is chosen to approximate the ratio of the top surface drag on the material, as can be developed by the latter minimum hugging pressure, and the total drag exerted on the material. Adherence to the design criteria 1, 2, and 5 assures that the former minimum hugging load requirement is fulfilled when designing specifically for the latter.
A very sensitive area in the design of the S-shape conveyor is where the curvature reverses instantaneously. This is shown on detail-B of Figure 24. Theoretically, no hugging pressure is lost if the curvature reversal is, in fact, instantaneous. Practically, however, an instantaneous curvature reversal is not possible, as is shown in Figure 24. A straight profile must exist between the idlers to either side of the reversal point when the belt cross section is fully loaded. The length of this straight portion is increased when the conveyor is only partially loaded. Theoretically, no hugging pressure, due to belt tension, exists on a straight conveyor profile (see equation 5). If the adjacent idlers to either side of the reversal point are sufficiently close, secondary resistance to prying the belts apart is instantaneous and increases in response to the increase in spreading distance between the two belts. Such response might make it possible to convey the material through this region without backsliding and to hold the stationary material when a shutdown occurs with a loaded belt. The interplay of forces is extremely complicated, and accurate assessment requires a sophisticated analytical model.
Several preliminary design trails were performed for a conveying rate of 6,000 tons per hour and simultaneous adaptability to mine slope angles from 40 to 50 degrees. Utilizing 84-in. wide standard multi-ply, nylon-by-nylon fabric belts, supported by 20-degree troughing idlers, resulted in a radius of vertical curvature at the bottom portion of the conveyor profile of 60 feet and a maximum elevating height of 248 feet. The elevating height was limited by the geometry of the profile when adapted to a 50-degree mine slope angle and not by the belt tension. The high hugging pressure requirement on the top belt in the area of the inflection point dictates a short radius of curvature. The combined radial force exerted on the curve by the two belts requires extremely close spacing of the support idlers to either side of the inflection point. Moving the inflection point further upward on the profile is not possible because the required idler spacing for an 84-in. belt would not be physically possible with standard idlers. Higher elevating heights could be achieved if the mine slope is decreased and or the capacity and belt width are decreased. The latter is because the idler load rating for a CEMA E7 idler is the same for the different belt widths.
This conveyor has better geometric conformance to the mine slope than does the Loop Belt, but it still deviates appreciably from the mine face. Such deviation means that the structural support must consist of a long span truss to follow the conveyor profile and react on the mine face only at the two end points. Access from the face to the conveyor would exist only at these end points. Such a unit would be self-contained in that the belt tensions and loads would be taken through the truss and would not require external anchorage. A truss to fulfill these requirements would be very heavy. Fabrication and erection would also be more difficult than for a truss which is essentially straight between the two end points and follows the mine face closely.
THE SNAKE SANDWICH CONVEYOR
The Snake sandwich conveyor is illustrated in Figure 25. If the S-Shape conveyor is a product of evolution, then the Snake
sandwich conveyor is the next step in the evolutionary process.
It, too, is a category (e) solution and requires the use of nylon fabric multi-ply belts.
Like the S-Shape conveyor, the Snake sandwich conveyor requires the ability to predict the minimum operating belt tensions along the profile in order to determine the profile geometry. It is, thus, greatly dependent on the use of automatic take-up arrangements at the tail end and a load sharing drive arrangement at the head end, as previously described.
The major difference is the introduction of many inflection or curvature reversal points on the conveyor profile, which allows for a geometry which is much better suited to the mine face geometry. The profile inflection points are also at lower conveying angles and, therefore, require less hugging pressure in this region.
The radius of curvature is, again, shortest at the bottom where the belt tension is lowest and increases with height corresponding to the increasing belt tension. The radius of curvature is continuously changing in order to provide only the required hugging load and, thus, prolong the life of the belt and reduce the idler requirements. Like the S-Shape conveyor. the Snake sandwich conveyor essentially dictates its own geometry and, thus, requires a trial-and-error approach; but there is more freedom in determining the geometry as inflection points are introduced when convenient.
Many preliminary design trials were performed for a conveying rate of 6,000 tons per hour and simultaneous adaptability to mine slope angles from 40 to 50 degrees. As with the S-Shape conveyor, two 84-in. standard multi-ply nylon-by-nylon fabric belts were carried on 20 degree troughing idlers. The resulting radius of vertical curvature at the bottom portion of the conveyor profile was 60 feet as in the S-Shape, but an elevating height of 330 feet, for the 50-degree mine slope, was achieved with standard cataloged belts. The elevating height was limited by the working tension rating of the two belts, with due regard to possible intermittent overload as a result of the inherent tolerance in the control system of the load sharing drive arrangement. The most significant result is that a reasonable profile geometry could be enveloped within a straight box truss of 15-foot depth. Such a truss depth is consistent with economical truss design for such an application (when considering the trade-offs between the additional cost of the intermediate support portals and foundations versus the cost savings in the truss itself) and would result in a three-point support for the design requirements stated. The simple geometry, reduced weight of the truss and its proximity to the mine face facilitate fabrication and erection and allow for easy access, as desired, from any mine bench level spanned by the structure. A conveyor unit of the specified lift and slope would be self-contained as was described in the S-Shape conveyor.
INTERMEDIATE DRIVES
Two types of intermediate drives were investigated for possible incorporation into the sandwich conveyor designs discussed. The first utilizes a driving belt which pulls the carrying belt over a required development length, as illustrated in Figure 26; while the second consists of pairs of driven tires which squeeze the belt at the edges, and transmit a tractive force to the belt, as illustrated in Figure 27. Both types of intermediate drives act to relieve the tension of the carrying belt forward of the drive so that the elevating height or length of a conveyor need not be limited, even for belts of relatively low strength. The Snake sandwich conveyor, as just described, presented an opportunity to apply either system in an attempt to achieve a single-run conveyor. The Snake sandwich conveyor appears to lend itself especially well to the drive belt-type intermediate drive, as shown in Figure 28, and presents a good opportunity to demonstrate the characteristics of both systems.
The drive belt-type intermediate drive depends on the normal lineal force exerted by the carrying belt on the drive belt. Based on a coefficient of friction of .5, the frictional development length can be determined. The Snake sandwich conveyor lends itself especially well to this type of intermediate drive because there is always a substantial radial pressure exerted by the belt sandwich along the curve, even when the material load is absent. Also, the geometry is convenient because it allows the drive belt to depart on a tangent at the inflection points of the conveyor profile without interfering with the carrying idlers. Driving the top belt in this manner did not prove efficient as the introduction of the conveyed material load caused substantial relief of the radial pressure along the curve, thus reducing the amount of frictional development.
Preliminary design for a 50-degree mine slope angle required a contact length between the carrying belt and drive belt in excess of 50 pct of the conveyor length. This requires a drive belt at every curve which supports the bottom belt. Many problems are inherent in a system of this type and they are listed below:
- Intermediate drives would be spaced at height intervals of approximately 50 feet. A 300-foot lift would require six drive units. The probability of breakdown is increased with the number of drives.
- Failure to drive the top belt requires a small increase in total hugging pressure when compared to the original design of the Snake sandwich conveyor in which both belts were driven. The increase is small because substantial edge distance exists and frictional drag along the edges acts to carry the middle portion of the belt.
- The introduction of the tensioned drive belt adds to the radial load to be supported on the carrying idlers and, therefore, requires more closely spaced idlers.
- The radius of curvature must decrease along the driven curve. As the tension on the carrying belt is relieved by the drive belt, the radius of curvature must decrease if the prescribed radial load is to be maintained. The hugging pressure from the top belt increases inversely to the requirement because its tension is constant and the conveying angle decreases. The radial pressure from the drive belt increases exponentially.
- Determination of the curved profile between the driven curves is largely dependent on the ability to predict the tension of the carrying belt at any point on the profile when, it alone, is driven. This means that a control system is necessary to assure that all drives contribute equally or nearly equally to the total driving effort. The complexity of a control system to coordinate load sharing would increase with the number of drives.
Such considerations are overwhelming and the incorporation of such drive units is ruled out.
The rubber-tire intermediate drive utilizes a squeezing pressure which allows the development of traction on the belt. Effective coefficients of friction between the tire and belt surface were found to be from .44 to .55 for wet and dirty surfaces(12). The combined traction-to-squeezing-force ratio of .8, as indicated in Figure 27, is conservative. The inflection points on the Snake sandwich conveyor are logical locations for this type of drive, as shown in Figure 29.
B.F. Goodrich and Continental Conveyor Company(13.14) have developed and used rubber-tire immediate drives. Drive modules of up to 200 hp are in existence. The system utilizes a special belt which has extra wear cover at the edges and limits reinforcements, for steel cord belts, to these edges. Testing with such drives leads the Russians, to the conclusion that ordinary belts can be driven in this manner without prematurely wearing out the cover at the edges.
Incorporating 200-hp drive modules into the Snake sandwich conveyor design, as shown in Figure 29, would require such units at height intervals of approximately 30 feet for a conveying rate of 6,000 tons per hour, and approximately 50 feet for a conveying rate of 3,500 tons per hour. A bigger unit would be needed for the higher conveying rate.
Considerations 1 and 5, listed for the drive-belt-type intermediate drives, are equally applicable to the rubber-tire-type intermediate drive. This makes their incorporation impractical.
If a straight-run conveyor is used, as few as two intermediate drives, utilizing steel cord belts could achieve elevating heights in excess of 900 feet. This is illustrated in Figure 30. However, the penalty is substantial because the required contact length between the carrying and drive belts is approximately 100 pct of the conveyor length (because μ = .5 for material on belt as well as belt-on-belt), and additional hugging pressure must be applied specifically to reduce this contact length. In addition, extra costs are incurred for drive belts, pulleys and take-up arrangements. Applying the rubber-tire-type intermediate drives to a straight conveyor is not different from its application on the Snake sandwich conveyor.
For the reasons discussed, intermediate drives are not considered practical at the present stage of development.
SUMMARY
Conventional belt conveyors offer an economical method for transporting bulk materials at inclines approaching 18 degrees. Many applications warrant conveying at steeper angles to 90 degrees, to minimize the support structure and/or excavation. Of the many high-angle conveying concepts known, the Sandwich belt conveyor appears to be the best solution for conveying at high rates, above 1,000 TPH and achieving high lifts, up to 350 feet per module. By studying the operational characteristics of past and present sandwich belt conveyors, the authors have established the design constraints and developed criteria for future development.
On this basis, new evolutionary methods in Sandwich belt high-angle conveying are suggested for future development. One class of solution is based on added mechanical or pneumatic pressure to the cover belt to develop the necessary friction between the material and the belt surfaces. The second class is based on hugging pressure derived by snaking the conveyor profile to exploit the inherent belt tension in any elevating conveyor. The use of intermediate drives in sandwich belt conveyors was investigated, but the many complications make such drives impractical with today’s technology.
