BELT CONVEYORS - DESIGN, OPERATION AND OPTIMIZATION
CONVEYOR DESIGN AND DESIGN STANDARDS
P. Staples Pr.Eng BSc. MSAIME Managing Director Conveyor Knowledge and Information Technology (Pty)Ltd (CKIT)
INDEX
1. 2. 3. 4.
INTRODUCTION JUSTIFICATION FOR A STANDARD PRESENT DESIGN STANDARDS PROPOSED STANDARD FORMAT 4.1 4.2 4.3 Selection 4.4 4.5 Drive Standards
5.
Power Pulley of
Belt Idler
and and Width
and
Tension Shafts Velocity Standards
CONCLUSION
SUMMARY This paper has been prepared with the intention of highlighting the problems faced by design engineers who are forced to undertake the design of belt conveyor systems using a multitude of design standards which have not been brought into line with modern technological advancements. To overcome some of these problems, a basic outline of a universal standard has been proposed, which can easily be adapted to suit individual needs, without reducing the efficiency of the designer and his team. 1. INTRODUCTION The design of belt conveyor systems has been one of the most common occurrences in the South African mining field for over one hundred years. Conveyors are seen on virtually all mining installations, and are the biggest problem for the plant maintenance engineer, being the cause of most plant shutdowns. Why do belt conveyors cause such problems? It must be remembered that mining houses usually have a set of design standards to conform to; standards which are claimed have been developed over many years to suit their own needs in the materials handling field. However, as I can understand the need for some aspects of a standard, others completely baffle me. It appears that having spent a great deal of time over certain requirements of a design standard, many of the fundamentals to which I am referring are of course the effects of overpowering on the whole conveyor system. Also, we know that to convey material from one point to another requires a specific amount of power using a belt designed to withstand a definite tension, so why is it that if a conveyor design problem is set to a number of designers, they will come up with many variations on a solution, even using the same design specification. This of course comes down to the interpretation of, and the familiarity with the standard to be used. Basically I am suggesting that the standards as available to-day, leave a lot to be desired from the point of view of completeness, and ease of application.
2. JUSTIFICATION FOR A STANDARD Do we need a standard at all? and if so, what form should it take? To answer this question let us look at a typical design office set up. On any project there are three key categories of staff, the designers, his draughtsmen and a group of peripheral staff, (planners, buyers, structural, civil and electrical engineers). Thus we have a set up which looks as follows:Figure 1. Typical project Engineering Flow Sheets
The designer is given a basic specification which will include material type and quantity to be conveyed from A to B. This he must transform into drawings for manufacture and fabrication, design data for civil, electrical and structural engineers, bills of quantities for buyers and activity networks for planners. With the exception of the planning information which is only really relevant for the construction phase of the project, the designer has a problem which he will find very difficult to overcome, and that is to supply all the necessary information to each discipline on the project when they require it. Therefore having obtained a scope of work from the client in question, the designer has to quickly produce the design data, but before he is able to proceed he must obtain information from his drawing office relating to the layout of the conveyors in the system. Now the problems begin: Prior to undertaking any calculations whatsoever the designer must check the specifications to which he must conform. As virtually all clients have their own opinion on the subject of conveyor design, we can rest assured there will be some form of client input, whether it be a two volume manuscript or simply an, 'All drives shall ........' document. The designer is confronted with conforming to the said specification, but much worse, he must ensure that his drawing office staff are aware that there is a specification to work to. Consider that the previous week they may have been working on another project and had to conform to a completely different specification. What does the designer do? Does he circulate multiple copies to his drawing office with the instruction that it must be read prior to any work being started. If so, he will possibly not meet his deadline on the supply of data to the peripheral disciplines. Does he try to check that his draughtsmen conform by 'looking over their shoulders' from time to time (which is the way mistakes are guaranteed to occur). Alternatively does he instruct his drawing office that there is a specification to work to and that it is lying around somewhere and to 'please check it if you are not to sure of how to proceed'. In all the offices in which I have worked, the last two solutions have been applied, with the result that, almost without exception, the experienced draughtsmen who know how to make a system work will continue with very little reference to the said specification. The problem may be that on this project 'the pulleys are much bigger, the take-up length must be selected using an ill defined formula and basically we don't know how to design a conveyor anymore. If this problem is caught early enough we only have to change a quantity of drawings and are then back on the right road. However you can be sure that in practice it will be too late, and the designer has to go to the client and ask for a concession because he is not able to conform to the specification, and to make any changes to the drawings now will put him way behind schedule. Furthermore before the client will accept deviations to the proposed format, every avenue must be explored, and a report on the deviation prepared. The designer is now behind whether he likes it or not and to make up time he must neglect the one function which completes the total conveyor design, that of secondary design. By secondary design, I mean the design which comes after the conceptual or general arrangement layouts are complete. This is the design of the chutes, the location of bearings, the belt cleaning system to be employed and the access for maintenance. This is left to a draughtsman without any engineering support. However, the secondary design usually encompasses the major problems of belt conveyor system design. These are areas with very little coverage in specifications, with comments such as, 'all conveyors will have pulleys at terminal points', being the limit to such specifications. I pose the question again, do we need a design standard? Those who agree with the scenario I have set will probably say, 'Allow the designer the freedom to do the job'. However 1 feel that a standard is essential. There are very few specialist conveyor designers and thus some form of guidance must be given. However there should
be only one standard, with one basic set of parameters and which can cater for the needs of every mining and process plant application. Without lessening the efficiency of the designer and his team such a standard will facilitate the efficiency the overcoming of the problems occurring in secondary design. We know this has been tried repeatedly in the past, but always in isolation from the main stream of design and usually with the statement, 'but it caters for our own individual needs', as justification. Having been confronted with conveyor design standards for a number of years, I have still to find a true specialist need, I know that some clients require less capacity on a belt, others require larger pulleys and thicker belts, requiring the use of complicated formula to arrive at a solution, but this can not be justification for devising completely individual specifications, which could more suitably be covered in a single paragraph of a comprehensive specification. 3. PRESENT DESIGN STANDARDS Let us look at at the Conveyor design standards available, and in particular the four most commonly used, C.E.M.A., GOOD YEAR, ISCOR and A.A.C. If we consider the power and tension variation predicted by using these systems, as in Table 1, we see quite a wide range of possibilities. The reason for this is in the selection of the rolling resistance factor, (coefficient of friction, resistance to flexure or other commonly used terms) which varies between 0,016 and 0,035 as used in the above standards. Table 1 Power and Tension calculations. 1(a) based on belt capacity of 500tons per hour, belt width of 900mm and a belt velocity of 2,2m/sec. Length
m
Lift
m
C.E.M.A.
GOOD YEAR
ISCOR
A.A.C.
Power
Tesn
Power
Tesn
Power
Tesn
Power
Tesn
kW
kN
kW
kN
kW
kN
kW
kN
30
0
6
9
15
16
16
19
12
18
200
60
101
65
99
64
104
66
102
66
1000
0
81
40
89
43
113
54
104
50
1000
40
132
72
143
77
167
88
158
84
1(b) based on belt capacity of 2000tons per hour, belt width of 1500mm and a belt velocity of 3m/sec. Length
Lift
C.E.M.A.
GOOD YEAR
ISCOR
A.A.C.
Power
Tesn
Power
Tesn
Power
Tesn
Power
Tesn
m
m
kW
kN
kW
kN
kW
kN
kW
kN
30
0
18
22
36
41
38
42
37
42
200
60
378
167
380
168
403
176
391
172
1000
0
221
84
262
98
349
127
315
116
1000
40
439
174
479
188
567
217
533
206
On the shorter systems this difference is quite insignificant, except that the belt length factor plays an important part. However on the now common large overland type systems, these variations are unsatisfactory to say the least. Are we able or prepared to accept such variations? Able, I will say yes, provided we take cognisance of the effects of overpowering. However I am not convinced we should be prepared to accept these variations, apart from the overpowering factor there are purely economic considerations to account for. This point is very noticeable when one becomes involved in economic evaluations (feasibility studies) of various alternative solutions to a specific materials handling problem. For instance, how competitive would a pneumatic conveying system or cable belt system be if designed to similar sets of standards as the conveyor. However as these standards are as yet, not
available, the manufacturer of competitive systems has far reaching advantages over the conveyor manufacturers. I am not for one moment suggesting that the competitive systems are under designed, simply that the designer is not limited to designing within a conservative specification. Too often we see examples of conveyor systems feeding process plants, where to conform to specification the whole conveyor network is designed for a large amount of excess capacity. However, this philosophy is not transferred to the related equipment in the rest of the plant. 4. PROPOSED STANDARD FORMAT 4.1 Power and Tension With power and tension calculations there exists the possibility for a combination of all four of the above standards by utilizing a single friction factor for the shorter belts, but eliminating the belt length factor which can easily be compensated for with the overrating factor of the motor. In progressing to the longer conveyors this factor could be variable, as advocated by C.E.M.A., only now be simply a function of belt length and capacity. Then we could use a simplified formula as follows:-
Power (kW) =
9.81 x L.V((kX+kY(Wm+Wb)+,015Wb)+ (H.Wm)) 1000
Where
• • • • • • •
L = Horizontal pulley centers (m)
•
kY = Resistance of the belt of flexure as it moves over the Idlers, and can be considered to be the same as the friction factors given in all the specifications.
H = Vertical pulley centers (m) V = Belt velocity (m/sec.) Wm = Mass of material per meter run (kg) Wb = Mass of belt per metre run (kg) 0,015 = Return belt resistance kX = Belt slide and Idler rotational resistance and can kX = 0,00068(Wm+Wb)+0,022(rotating mass of the Idler per meter) (kg/m)
be
obtained
from:-
Typical values of kY are given in table 2 below. Table 2. Selection of kY factor based on Belt length, lift and capacity. Length
Lift
kY
kY
kY
kY
m
m
500t/hr.
1000t/hr.
2000t/hr.
3000t/hr.
100
20
0,035
0,030
0,026
0,022
200
20
0,032
0,026
0,022
0,020
200
40
0,030
0,022
0,020
0,020
400
20
0,030
0,022
0,020
0,020
400
40
0,026
0,020
0,020
0,020
800
40
0,022
0,020
0,020
0,020
1000
40
0,020
0,020
0,020
0,020
To enable the client to maintain control of the outcome of the calculation, it is necessary only to specify the kY factor to be used in a simple addendum to the main specification. Belt tension calculation can be kept straightforward, provided the designer starts by considering the minimum belt tensions, at both the drive and tail pulleys, by using the following formulae :-
Tmin = 4,2x9,81/1000 si(Wb+Wm) kN Where
Si
4,2 = = Idler spacing,m
Factor
based
on
a
3%
belt
sag.
and Tslack side = Teffective / e -1 Where T effective is the installed drive effective tension and not the effective tension computed from the above power formula. The one problem that is encountered is in the selection of a coefficient of friction for the drive pulley. A standard such as given In Table 3 could be used. Table 3 Coefficient of Friction for Drive Pulleys. Type of Take Up Automatic
Manual
Plant Description
Conveyor Construction
Lagged
Unlagged
Lagged
Unlagged
Wet
Covered Uncovered
0,25 0,20
0,10 0,10
0,20 0,20
0,10 0,10
Semi-wet
Covered Uncovered
0,30 0,25
0,20 0,15
0,25 0,22
0,18 0,13
Dry
Covered Uncovered
0,35 0,30
0,22 0,18
0,25 0,25
0,20 0,15
Table 3 has been compiled from empirical data such as that given in Table 4. It should be noted that these values are the limiting conditions (when the belt is on the point of slipping). The actual coefficients of friction developed between surfaces are, in practically all cases where slipping does not occur, in excess of those listed. Therefore, the convention of using these values does not reflect what actually occurs at the drive pulley. If one considers a drive pulley under operating conditions then the higher tensioned belt section is stretched more than on the lower tensioned section, thus the belt entering the positive drive will be traveling faster than when it leaves it. The elastic recovery of the belt occurs over only a part of the total angle of contact, and it is at this point, where creep takes place, that the driving is done, while making full use of the coefficient of friction. By applying the classic tension formula to the whole angle of wrap a fictitious coefficient of friction is being used Table 4. Recommended Drive Coefficient of Friction of Various Standards. Condition
C.E.M.A.
STEVENS ADAMSON
BRIDGESTONE
LINATEX
REMA TIP TOP
Bare pulley
0,25
0,35
0,20
---
---
Lagged
0,35
0,35
---
0,60
0,45
Dry Lagged
0,35
0,35
0,35
0,60
0,45
Wet Lagged
0,35
0,35
0,25
0,80
0,35
Wet & Dirty
0,35
0,35
0,20
0,40
0,25
The advantage of working from minimum drive tension back to the maximum drive tension, can be better explained if one looks at the design of pulleys and shafts. Over the years there has been a lot written about the design of a pulley shaft, with the aim of trying to eliminate the high failure rate and the cost associated with such failures. I feel that there are only two basic reasons for pulley failure, firstly the bad manufacturing procedures, and secondly, failure owing to an inability to calculate the minimum drive tension. The latter case of incorrect design
results in the counterweight mass having to be increased to overcome drive slip on startup, with the result that pulley shafts are subjected to excessive loads, producing eventual failure. By contrast, if the minimum drive tension is used as a design basis, we can overcome, failures in pulleys, caused by inaccurate design. Thus the maximum tension will be obtained from :Tmaximum = Tminimum+Teffective Where Teffective is computed from shaft power and not the installed power. Note that the formulae discussed above are applicable to 90% of the conveyor installations being designed today. However a little more analysis is required for some overland and complex systems. 4.2 Pulley and Shaft Standards There are presently two major standards used for pulley and shaft selection, these being the ISCOR and AAC systems. I know much has been written about the high degree of oversizing adopted by both standards, but I feel that as the pulley is one of the least expensive components in a conveyor installation, we should not be over concerned on the point. Efforts should rather be directed at reducing the amount of variations there are in the selection of face width and bearing centers. At the moment both ISCOR and AAC have two sizes per belt width, all different. This should be reduced to a single size per belt width, and this size should be as big as possible to allow easy access and hence reduce the damage to conveyor belts. A standard along the lines of table 5 based on the ISCOR specification would be the most acceptable. Pulley and shaft diameters should be kept to a minimum of two per conveyor, with as much standardization as possible being employed on the whole conveyor system. The selection of pulleys and shafts could be from a table similar to that shown as Table 6. Table 5. Pulley Face Width and Bearing Centers Belt width mm
Face width mm
Bearing center mm
450
550
890
600
700
1140
750
900
1370
900
1050
1520
1050
1200
1670
1200
1350
1850
1350
1500
2000
1500
1700
2300
1800
2000
2630
2100
2300
2930
4.3 Selection of Belt Width and Velocity The selection of belt width and velocity is probably the most frustrating of problems facing the designer. There are a variety of factors being used, factors such as :- the belt width must be three times the maximum lump size, the belt width must be such that the system can cater for 66% excess capacity, and if a tripper is used the factors must be increased by a further 30% etc. This type of factor forms the basis for most standards in use to-day, and these could therefore be rationalized into a single more acceptable standard to make the designer's task easier. The first necessary step is the removal of the age old belt speed restrictions, after all speeds in excess of 4m/sec are now quite common.
I am not advocating that the highest possible belt speed be used for all installations; I simply suggest that belt speeds should not be selected only on the basis of past experience, but on the basis of belt length, transfer point and economic considerations. I feel that to use the criterion I have set out will automatically result in the selection of the most suitable belt width and speed. My reasoning here is that, for inplant installations belt widths and speeds are almost always selected on the basis of standardization and possible transfer point problems. By contrast, the larger overland systems are selected on the basis of capital costs and the associated operating and maintenance costs, because as belt speeds increase operating and maintenance costs usually follow suit. Consider the suggested methods of selecting a belt width and speed. Firstly, the amount of material on a belt must be related to the expected transfer point problems. A flat feed point fed by a controlled system will be far easier to design than an inclined feed point fed from a crusher, where surges are very common. Therefore to suggest a similar standard for both applications is not practical. We often are told that conveyors should not be fed at angles of 8° incline feed points and very tight vertical curves, with the result that the feed point stays clean, but at the curve the belt has lifted causing spillage. I would like to suggest that a belt can be easily fed at angles of up to 16°, provided the belt width and speed are correctly selected. It may be necessary to install belts with thicker covers, but this can form the basis for a better design. Thus the type of standard that could be used is shown in Table 7. Table 7 Implant Conveyor Load Factors Loading Point Type
Feed Type
Overload Factor
Horizontal
Uniform
1,20
Horizontal
Surge
1,50
Incline
Uniform
1,50
Horizontal
Surge
1,75
Tripper
.....
1,75
Shuttle
.....
1,50
The overload factor would be used to increase the design tonnage for selection purposes. For overland conveyors it is common to use horizontal loading points, and we are not confronted with the same problems. As mentioned earlier it is only necessary to consider the economics of the system, with the following limitations as given in Table 8. Table 8. Overland Conveyor Minimum Belt Widths and Maximum Speeds Terminal Pulley Centers (m)
Belt Width (mm)
Belt Speed (m/sec)
300 to 500
600
3,50
500 to 1000
750
3,50
over 1000
900
7,00
The overload factor used should always be a minimum of 1,2 times the design tonnage. 4.4 Idler Standards 4.4.1 Introduction The introduction of the SABS Idler specification will ensure a more uniform selection of idlers. As a result the choice of type and spacing for Idlers should be on a more scientific basis. The types of Idler to be used on conveyors are; transition, troughing, impact and return idlers. At this time there is no satisfactory training idler available so they should be avoided.
4.4.2 Troughing Idler Spacing Two types of troughing idler are used frequently, fixed and suspended roll. There is very little difference between the two, except the training characteristics and possible cost savings associated with the suspended roll. The question of idler spacing needs be considered more carefully. The restrictive standards as applied to-day do more harm than good to a conveyor system. Idlers are the highest maintenance cost item on a conveyor installation and the biggest cause of belt damage, therefore 'the fewer the better. Idler spacing must be selected on the grounds of available belt tension, fatigue life of the idler bearings, and structural considerations. The upper spacing limit should be set at 2200mm. Account should be taken of four and five roll sets, but no significance can be attached to the claim that four and five roll idlers give better belt life. 4.5 Drive standards The standardization of drives is the key to most successful conveyor systems. The problem is however that some drives have to be drastically oversized to obtain some degree of conformity. By considering this point at an early stage in the design process. it is usually possible to overcome the problem, therefore simple cost analyses of all the possible solutions can quickly decide on the drive sizes to be adopted. Also it is at this point in time when a final selection of belts can be carried out, because there is often scope to change belt speeds to the required degree of standardization, and we should not be afraid to to this. 5. Conclusion To conclude I would like to reiterate the need for a single design standard, which could be applied to any conveyor installation. However, this standard must be such that it allows the client a small amount of individuality and flexibility. The design system as outlined in this paper can offer this flexability, by allowing the client the freedom to select the kY factor, the drive coefficient of friction and the load factor for selecting the belt width and speed. Coupled with this we can have a very efficient system especially if it is adapted to computerised calculation techniques. I know to-day that many such design programs are available, but because of the variations in standards that must be incorporated, their credibility is unjustly made suspect, forcing the designer to revert to the longwinded number crunching exercises which obviously reduce his effectiveness in the drawing office. BELT
PULLEY
HEAVY DUTY
MEDIUM DUTY
LIGHT DUTY
BELT TYPE
450
300 400 500 630
90 100 125 140
75 75 100 110
75 75 100 110
50 50 75 90
75 90
50 75
200 250 630 800
18 22 56 72
600
400 500 630 710
110 125 140 160
90 100 110 125
90 100 110 125
75 75 90 100
75 90 100 110
50 75 90 90
250 630 800 1250
30 75 95 150
750
400 500 630 710 800
125 140 160 180 200
100 110 125 140 160
100 125 140 160 180
75 100 110 125 140
75 90 110 125 140
50 75 90 100 110
250 630 800 ST500 1250 ST630 1250 ST1250
35 95 120 190 280
900
400 500 630 710
140 160 180 200
110 125 140 160
110 140 160 180
110 140 125 140
90 110 125 140
90 110 100 110
250 630 800 1250
45 110 145 225
MAXIMUM SHAFT SHAFT BEARING SHAFT BEARING SHAFT BEARING PLY STEEL LOAD CLASS CORE WIDTH DIA.D DIA.d DIA.d1 DIA.d DIA.d1 DIA.d DIA.d1 kN
ST500 ST630
800
220
180
200
160
160
All DIMENSIONS MAXIMUM LOAD FIGURE = PERMISSIBLE LOAD ON HEAVY DUTY = 100% MEDIUM DUTY = 60% LIGHT DUTY = 30% DENOTES RATING BASED UPON FOR ALLOWABLE LOAD ON BEARING SEE BEARING RATING TABLES
125
1250 ST1250
340
IN PULLEY
MILLIMETRES = TWICE BELT TENSION MAXIMUM LOAD MAXIMUM LOAD MAXIMUM LOAD STEEL CORE BELT
BELT RATING CHART, Table 6a BELT
PULLEY
HEAVY DUTY
MEDIUM DUTY
LIGHT DUTY
BELT TYPE
SHAFT BEARING SHAFT BEARING SHAFT BEARING PLY STEEL WIDTH DIA. D DIA. d DIA. d1 DIA. d DIA. d1 DIA. d DIA. d1 CLASS CORE
MAXIMUM SHAFT LOAD kN
1050
500 630 710 800 1000 1250
180 200 220 240 250 360
140 160 180 200 220 340
140 160 180 200 220 260
110 125 140 160 180 220
110 110 125 140 160 180
75 90 100 110 125 140
630 800 ST500 1250 ST630 1250 ST1250 1600 ST1600 2000 ST3150
130 170 260 395 505 985
1200
500 630 710 800 1000 1250
180 200 220 240 260 360
140 160 180 200 220 340
140 160 180 200 220 300
110 125 140 160 180 260
110 125 140 160 180 220
90 100 110 125 140 180
630 800 ST500 1250 ST630 1250 ST1250 1600 ST1600 2000 ST3150
150 190 300 450 575 1130
1350
500 630 710 800 1000 1250
200 220 240 280 300 360
160 180 200 240 260 320
180 200 220 240 260 300
140 160 180 200 220 260
140 160 180 200 220 240
110 125 140 160 180 200
630 800 ST500 1250 ST630 1250 ST1250 1600 ST1600 2000 ST3150
170 215 340 505 650 1270
1500
630 710 800 1000 1250 1400
240 280 300 320 360 400
200 240 260 280 320 380
200 220 240 260 280 320
160 180 200 220 240 280
140 160 180 200 220 240
110 125 140 160 180 200
800 1250 1250 1600 2000 2500
ST500 ST630 ST1250 ST1600 ST3150 ST4000
240 375 560 720 1400 1800
1800
710 800 1000 1250 1400 1500
300 320 340 380 410 430
260 280 300 340 380 400
260 280 300 320 340 360
220 240 260 280 300 320
180 200 240 260 380 300
160 180 260 220 240 260
1250 1250 1600 2000 2500
ST630 ST1250 ST1600 ST3150 ST4000 ST5000
450 670 865 1700 2100 2700
2100
710 800 1000 1250 1400 1500
300 320 340 380 410 430
260 280 300 340 380 400
260 280 300 320 340 380
220 240 260 280 300 340
180 200 240 260 280 300
160 180 200 220 240 280
1250 1250 1600 2000 2500
ST630 ST1250 ST1600 ST3150 ST4000 ST5000
525 790 1010 1900 2500 3100
SEE GENERAL NOTES ON SHEET 1
BELT CONVEYOR DRIVES - A CONSIDERATION OF SOME DESIGN ASPECTS
J.H. Rall Pr.Eng., BSc Eng., MSAIME Hansen Transmissions (Pty) ltd P. Staples Pr.Eng BSc, MSAIME Managing Director Conveyor Knowledge and Information Technology (Pty)Ltd (CKIT)
Summary: This is a short review of part of the link between the electric power grid and flat rubber covered belts used for transporting large volumes of granular material. It is concerned with high volume material conveying and not with special cases such as feed or metering conveyors, steep inclined conveyors etc. It considers mainly the speed reducer between motor coupling and drive pulley, ratings, bearing life, service factors, stopping and antirunback devices. 1. General: The ever increasing rate of consumption of earth's raw materials has brought with it a need for faster movement of these materials from the point of extraction to the point of process or usage and transporting these materials through the process plant and disposing of the waste in the shortest possible time. Many methods of material handling are employed from wheel barrows to dump trucks or shuttle cars, to pneumatic ducts carrying pulverised particles in an air stream. In this line of movement, belt conveyors play a very important part in the reliable carrying of material over long distances at a competitive cost. Each method of material conveying has its advantages and disadvantages. One of the problems with belt conveyors is that soft friable material can be degraded, particularly in loading and unloading. If the maintenance of lump size is important, this can present difficulties on a complicated conveyor system. Conveyor systems have become larger and more complex and drive systems have also been going through a process of evolution and will continue to do so. Bigger belts require more power and has brought the need for larger individual drives as well as multiple drives such as 4 drives of 1000 kW each on one belt. Shaft mounting of the complete drive unit is another change which has brought with it the requirement for more compact and lighter drive units. This tends to favor a right angle drive configuration with the motor next to the belt and hardened gears to reduce the dimensions and mass of the drive. 2. Drive Ratio and Belt speeds: Depending on the quantity, size, distance and characteristics of the material to be conveyed, the absorbed power, width, tensile requirements and top cover thickness of the belt will be decided. Large volume conveyor belts run in the range of 2 to 6 metre/second and the allowable bend radius of the belt determines pulley diameters which for large belts is of the order of 0,8 to 1,5 m giving pulley speeds between 50 and 125 rpm. Assuming that 4 pole motors are used, this gives a reduction ratio required between 12:1 and 30:1. Most modern gear manufacturers do not use a higher ratio per stage than 5:1, which means that speed reducers will be either 2 or 3 stage reduction. (Except for small powers where worm reducers, or torque arms and V belt drives may be used). There is a misconception that one can reduce the cost of the gear-speed reducer by using a 6 or 8 pole motor, but even an 8 pole motor on the higher speeds would require a reduction above 6:1 and a 2 stage unit would still be required. The bulk of the cost of a gearbox is related to the low speed shaft torque and therefore having determined this, there is generally no economic advantage at all in using anything but a 4 pole motor. The motor manufacturers, because of size and volume, generally supply 4 pole motors at the lowest price, and as a rule therefore, a 4 pole motor is the best choice with a gearbox of the appropriate ratio to arrive at the desired conveyor shaft speed. Where ball and roller bearings are used in the electric motors some manufacturers prefer 6 pole or even 8 pole speeds for motors over 1000 kW. 3.1 Choice of Single or Multiple Drive.
Having calculated the power required to drive the belt and having considered the belt tension and angle of contact, a decision can be taken on whether the belt should be fitted with single or multiple drive. This decision is often influenced by other equipment installed in a plant and multiples of other smaller drives are often used. Drive size may also be determined by the nearest standard motor available. Where a drive point is situated some considerable distance from the main power source, a long cable may be involved to supply electric power to the drive. In this case, the cable size and cost of transformers may play an important part in the selection of number and size of motors used. With direct on line starting, the peak current the motor will draw is likely to be of the order of 6 x full load current and the combination of running current of a motor or group of motors with the starting current of the last motor to start will have a strong influence on the drive choice. 3.2 Method of Low Speed Shaft Connections. Drive from the reducer to the belt pulley shaft is either by "flexible" coupling from a drive pack mounted on a foundation next to the structure or by shaft mounted drive unit hanging on the pulley shaft. When shaft mounted, the drive unit can be either hollow shaft, driving through a friction locking element or solid shaft attached by a rigid coupling to the conveyor shaft. Some typical attachments are shown in the sketches. (See appendix A). 4.1. Choice of Starting method, drive size and Protection: During start up of conveyor belts, a considerable mass is usually involved which requires acceleration, and to reduce the length of time that the motor draws starting current, a "slip coupling" is fitted between the drive motor and speed reducer. Alternatively, slip ring motors are used to achieve a quick but gentle start up with control of the peak current. On small belts below 10kW direct on line starting directly coupled is quite normal and on belts, say below 100 kW D.O.L. starting with "slip" couplings is most common, and probably the simplest and most cost effective. On larger drives with power at its present continuously increasing cost, slip ring motors may be attractive, due to the prevention of the peak, particularly where maximum demand plays a part in the electricity tariff. There is a multitude of slip couplings on the market for use with D.O.L. start motors, but for larger belts the majority in use are liquid type couplings, either straight traction or traction with delayed fill or controllable fill (scoop type). A fluid coupling will always "slip" a small amount and will help multiple drives to share load, provided the coupling "fill" has been carefully adjusted. As a rule, each coupling has a slightly different characteristic and if adjusted to share load correctly under full load conditions will more than likely not share properly under light load conditions. Motor characteristics also vary a little and can also contribute to poor load sharing on multiple drives. Drive systems commonly go up to 4 motors per belt, but rarely more. 4.2 Motor Starting: (D.O.L. with fluid coupling). On multiple drives accepted practice is to start the motor on the secondary drive first and say 3-5 seconds later one of the primary drive motors and then the next primary motor say 5-10 seconds later. In practice, however, the observed starting procedures and delay times vary a great deal. A very common sight is to see the secondary drive motor running and due to conservative coupling selection, the starting current drops somewhat, but starts rising quickly again due to the delayed fill coupling increasing it's slip torque, while the belt remains stationary. If a primary drive is started at the correct time before the coupling torque has increased too far, the belt is brought into motion much quicker with a lower overall current. The relative slip of the coupling affects it's torque and so the motor current, and the sooner the belt can be moved, the sooner the current peaks will drop. In the case of scoop controlled fluid couplings all motors are started in quick succession and then all couplings filled slowly. A similar procedure is followed with slip ring motors and these two methods are undoubtedly the kindest to the belt, pulley and lagging etc. The choice of scoop type fluid couplings or slip ring motors is likely to lead to the use of smaller motors with safety and possible savings on switch gear maintenance. 4.3 Other Drive Methods: Where a variable conveying rate is required, D.C. drives can be used as well as squirrel cage motors with frequency control. Another method is by hydro-static drive; again ideal for variable speed, but overall running and maintenance costs on big powers are likely to be higher than a fluid coupling drive, with S.C. motor or slip ring motor.
4.4 Belt Protection: Belt protection against overload and stalling is commonly done by a centrifugal switch driven by a roller on the underside of the belt. This, however, is not very sensitive and more sophisticated methods are now used. One method consists of fitting a pulse generator to the drive gearbox low speed shaft and similar pulse generator to a roller driven by the belt. A monitoring unit compares the pulse frequencies continuously and if they go outside set limits, an alarm is given or the belt stopped. On start up, belt slip can be kept to a minimum by using the monitoring unit to control the start up on slipring motors. The monitoring unit can control the rotor resistance and so the starting torque. Alternatively, the "fill" of a variable fill fluid coupling can be controlled by the comparitor. 5. Stopping a Belt Conveyor 5.1 Forward stopping. As a general rule, friction will reduce the normal forward speed of the belt and load and bring it to rest in a relatively short time. The allowable time for stopping depends mainly on the discharge end conditions. Where one belt feeds onto another, tripping conditions, transfer bunker size and belt layout may indicate a need for a belt to be slowed down by other methods than normal friction. On a downhill section of a conveyor discharging onto a level or uphill belt is generally the place where braking is required. If one belt runs on longer than the rest of a system of conveyors, bunkers or transfer chutes can be overloaded and may even be a hazard, but this is generally a very rare condition. Considering a conveyor layout as per sketch (see appendix B) fitted with a brake on the gearbox high speed shaft, it can be seen that when stopped under load the inertia of the load would tend to pick up the gravity tensioner and release the tension in the belt between drive and head pulley and may even go slack by the time the load comes to rest. The load assisted by the gravity take up would then accelerate the belt in the return direction while the drives are stationery with brakes on. When the belt between drive and head pulley becomes tight it has to retard the load in a fraction of the distance in which it had been accelerated. Under this sudden stopping of the load the drive is subjected to heavy shock. In this case, a brake is not only unnecessary, but highly undesirable. A brake fitted on a downhill belt drive would again release tension between drive and head pulley and pick up the take up, but this would tend to release tension on the driving (or stopping) pulleys and allow the belt to slip. This can damage the belt and pulley lagging and can also be dangerous. The proper method of stopping a belt like this is on the tail pulley or other pulley on the return belt after it leaves the take up. Another application where brakes are sometimes used is on belts running through a mobile stacker, to reduce the risk of the belt snagging; should the stacker be moved while the belt is stationary. Here again a brake would be fitted to the tail pulley and a holdback to the drive or head. 5.2 Reverse Stopping: Uphill conveyors normally only need an anti-runback device. This can be fitted to the head pulley or on one of the drive pulleys or to the gearbox. If fitted to the gearbox, the torque which the holdback is required to resist is reduced by the ratio of the gears between the holdback and drive pulley. (See appendix C). Considerable savings can therefore be made on the size of holdback fitted to the gearbox compared to that required on the head or drive pulley. However, on multiple drives there can be problems with load sharing. On normal running and start up, the driving load is shared between the various drives by either fluid couplings or slip-ring motors as previously mentioned. When being held by a holdback fitted to the gearbox, the system is torsionally much more rigid than in the drive direction and it is highly unlikely that more than one gearbox / backstop combination would hold the belt unless specific provision is made for load sharing. On a continuously inclined belt, (which is the worst,) at the moment of coming to rest, the friction load can be considered not to play any significant part in holding the belt back and the full driving torque has to be considered acting in the reverse direction. On a multiple drive in this situation, if the strength of one gearbox is sufficient to resist the runback torque, then a holdback on one gearbox will be adequate. Alternatively, more than one gearbox has to be fitted with a holdback and some means of load sharing provided. (See appendix D for holdback sketches). In many cases brakes are fitted together with holdbacks. If both the brake and holdback have each been selected for a torque somewhere near the driving torque, then clearly on runback condition with brake and holdback
working together, their combined torque is likely to be considerably higher than the gearbox rating. Returning to the sketch of the conveyor just discussed, the brakes will bring the drive pulleys and gearboxes to rest long before the load. Consider the gearbox high speed shaft fitted with brake drum and holdback both locked against reverse motion. When the belt suddenly applies reverse drive to the pulley, the holdback and braking torque has to be overcome and the brake drum accelerated through the gearbox as speed increases. In this situation, an incredible stress is placed on the drum, low speed coupling and gear-train. 6. Service Factors: (See appendix E for service factors recommended by AGMA). 6.1 Gear Rating: The use of a service factor is often regarded as a factor of ignorance by the end uses. It may well represent the ignorance of the drive suppliers of the true conditions. In broad terms, gears rated according to the AGMA method of calculation will transmit the rated load continuously, with a risk of tooth failure of less than 1% due to surface durability or strength provided that no shock loads are applied, no bearings fail, no material defects exist, load bearing is perfect and lubrication is perfect etc. In practice, a steady load is not very common, neither are bearings which do not wear or absolutely rigid housings, shafts etc. A service factor should therefore be selected with due consideration of the actual conditions and not in "ignorance from a table" 6.2 Load Conditions on Teeth and other Components: The loading and stress on any one gear tooth will vary from zero to maximum once per revolution tensile on one flank and compressive on the opposite as long as it runs and drives in one direction. If this same drive unit is used on a reversible drive, the variation of stress on a tooth is therefore increased as both flanks of the teeth are subjected to tensile and compressive stress and the range from maximum to minimum nearly doubled and the fatigue life of the teeth considerably reduced. Even under steady load conditions in both directions, the 1% chance of failure will no longer apply. Another potential problem area on reversible drives is keys and keyways. Keys invariably have a very small amount of side clearance which will move from one side of the key to the other on reversal. The effect this has will mainly depend on the amount of clearance and the shock which is applied. This is normally a maximum under start up when backlash between the teeth, coupling, elements etc., are likely to be open particularly on the reverse start up condition. Reversal of stress in drive components can also occur on units running only in one direction, for instance, where brakes are fitted to slow the belt down. The above for instance, applies equally to the low speed shaft couplings where a spring element on a reversing conveyor may fail in fatigue where an identical drive on a unidirectional conveyor shows no sign of fatigue. The choice of brakes on conveyor drives appear to be selected to suit the drive motor torque or more. On a crane, this may be a valid starting point for brake selection, but on a conveyor which may have up as well as downhill sections, the motor torque is likely to lead to the selection of much too large a brake. On the use of service factors, it is common practice to select a gearbox with a minimum service factor calculated on installed motor power. The motor power is often arrived at by doing very extensive and careful calculations of the power required to drive the belt which in simple terms will consist of power to overcome friction plus power to lift the load. In both cases, an absolute maximum can be taken and then a contingency factor added. The efficiency of the fluid coupling can then be estimated, to be extremely low and so also the gearbox efficiency. Apply a service factor on top of this, then select the next bigger standard motor and the drive motor will be able to cope with the most adverse freak condition with power to spare, but do a fair amount of damage to the belt, pulleys etc. The last thing one wants to do is select a drive unit too small, but it is essential to do all the calculations accurately and then apply a "contingency" factor only once. The consequences of selecting a gearbox or a low speed shaft coupling on drum shaft too large would produce an uneconomic solution, but would give a very long expected life. However, selecting too large a motor will have quite the opposite effect on the life of a belt,
pulleys, gearboxes etc., over an extended period of time. As an example, consider a conveyor drive calculated to require 70 kW to drive it and add a 50% contingency factor. indicating an absolute maximum 105 kW. The next bigger size of motor is say 132 kW. On start up the drive train may be subjected to a torque equivalent to 2½ x 132 Kw = 330 kW while originally one set out only needing 70 kW. On start up, the mechanical components of this drive will certainly receive a beating. Returning to the calculated rating of the drive unit; to ensure that one maintains the chance of failure at the correct level, the operating conditions have to be analysed and a S.F. selected to ensure that shock and other unforseen conditions do not go outside the factors taken into the calculation. It is accepted that the calculated rating of a gear can be exceeded for short periods without damage. For instance, it allows for a starting torque of twice the AGMA rating of the gear without ill effect. 6.3 Bearing Life: The normal methods of rating a ball or roller bearing in a gearbox is to assume a B10 life of 5000 hours and calculate the rated power of the gears for this bearing life which is then taken as the bearing rating. The kW rating for bearings will therefore normally be higher than the rating of the gears. The generalised relationship between the life of roller bearing, the operating load and rated capacity is expressed as:Life =
Constant
[ Rated load capacity
]
3,3
Operating speed [ Actual operating load ] When the bearing rating at a specific speed is calculated as stated above in terms of kW for 5000 hours B10 life and the average absorbed power of the driven machine is known, then the expected B10 bearing life will be:5000 * [
kW rating of bearing
[ Absorbed average power in kW
]
3,3
Hours
]
Statistically more than 90% of all bearings operating under these conditions will have a life at least equal to the calculated figure or the chance of failure is less than 10% in this period of running time. This is a statistical probability based on the assumption that the actual life will be normally distributed and the 10% failures include the possibility of a failure in the first minute of operation. likewise the 90% which will exceed this calculated life will include some which will have infinite life as the tails of the normal distribution curve stretch from zero to infinity. For other chances of failure, the bearing manufacturers give the following factors with which to multiply the B10 life to calculate the expected life for the corresponding chance of failure. Probability of failure 10% 5% Factor
1
4%
3%
2%
1%
0,62 0,53 0,44 0,33 0,21
The catalogue rating for a gear unit is the rating in kW of the weakest element in the gear unit. Assuming the buyer requests a S.F. of 1,5 on the absorbed power, and that a bearing is the weakest element, then the expected bearing life will be approximately 19000 hours for a 10% chance of failure or 4000 hours for 1% chance of failure. The relationship between load and life is exponential and a small change in load will have a significant influence on expected bearing life, therefore, it is essential that the average absorbed power of the belt be calculated. 7. Conclusion: To provide a first class solution to a materials conveying problem an important point is a good integrated balanced design of which the drive is one element which requires close and open collaboration between the conveyor designer and the drive supplier. To this end, may we suggest that when placing an enquiry for conveyor drives of say over 100 kW, that the following information be given:-
1. 2.
3. 4. 5. 6. 7.
A simple sketch of the conveyor layout. Calculated absorbed power in some detail such as:a. Power to overcome all friction losses. ........... ... kW b. Power to lift load. (If lift is involved). ........... ... kW c. Power to move load on level. (Where level and lift involved). __________kW d. a + b + c = Calculated normal total absorbed power. ========kW e. Contingency for overload and frequency expected (or chance of overload in %). ................kW f. Total maximum calculated absorbed power. ========kW g. Additional torque required to accelerate load ............ .... Nm If this is more than 75% of the equivalent torque Of (f), then it needs consideration in relation to the motor characteristics. h. For purposes of bearing life calculation, the expected average continuous absorbed power. Method of starting and expected start up time. Method of preventing runback. Size of motor (If not less that 'f' reason for choice). Expected service factor and reason for choice of S.F. Expected bearing life and acceptable chance of bearing failure.
The end user requires larger outputs for a specific capital costs coupled to low energy consumption, low maintenance costs and high reliability over a long operating life. One of the "elements in this "mix" is the speed reducer for which the optimum choice is at the present moment certain to be a helical gear unit with shafts running on ball or roller bearing with hardened teeth driven by an appropriate "mix" of electric motor and coupling External Shrink Double Taper. Appendix A1
Internal Friction Locking Elements
External Shrink Disc Single Taper Appendix A2
Internal Stretching Element Appendix A3
Solid Shaft Rigid Coupling Shaft Mounting for Gearbox
Conveyor Drive with Friction Brakes Appendix B
Normal Running
Full load trip Brakes stop belt and load carries on picking up gravity take up causing slack belt. If brakes are required "X" is the proper place to apply them. Gearbox Ratio 16:1 Then Torque on Backstop 1/16th. of L.S.S. Torque Appendix C
L.S. Gearset Ratio 4:1 Torque on Backstop on Intermediate Shaft ¼ of L.S.S. Torque
Roller and Ramp Type Holdback Appendix D1
Sprag Type Holdback
Centrifugal Release Sprag Type Holdback Appendix D2
A = Locking Direction B = Free Wheeling Direction Band Type Holdbacks Appendix D3
APPENDIX E WITH ACKNOWLEDGEMENT TO AGMA. AGMA STANDARD PRACTICE FOR ENCLOSED SPEED REDUCERS OR INCREASERS USING SPUR, HELICAL, HERRINGBONE AND SPIRAL BEVEL GEARS 4. Service Factors 4.1 Before a reducer can be selected for any given application, the equivalent horsepower is computed by multiplying the specified or actual horsepower by the service factor for the particular load classification for which the unit is to be used. It is necessary that the unit selected have a capacity equal to or in excess of this equivalent horsepower. The recommended service factors for various load classifications and duration of service are shown for several types of prime movers in Table 1. 4.2 Load classifications for various applications are given in Table 2. They are classified into three commonly recognized load classifications : Uniform, Moderate Shock, and Heavy Shock. 4.3 Service factors represent the normal relationship between gear design power rating and the continuous power requirements. Applications involving unusual or severe loading or requiring a high degree of dependability should be carefully reviewed with the manufacturer before a service factor is applied. 4.3.1 Applications with high-torque motors and motors for intermittent operations and applications where extreme repetitive shock occurs or where high-energy loads must be absorbed, as when stalling, require special consideration and are not covered by the service factors given in Table 1. Table 1 Service Factors
Prime Mover
Duration of Service
Electric Motor, Steam Turbine, or Hydraulic Motor
Occasional 1/2hr. per day Intermittent 3hrs. per day Over 3hrs up to and incl. 10hrs. per day Over 10 hrs. per day
Multi-Cylinder Internal Combustion Engine
Single Cylinder Internal Combustion Engine
Occasional 1/2hr. per day Intermittent 3hrs. per day Over 3hrs. up to and incl. 10hrs. per day Over 10hrs. per day Occasional 1/2hr. per day Intermittent 3hrs. per day Over 3hrs. up to and incl. 10hrs. per day Over 10hrs. per day
Driven Machine Load Classification Uniform Moderate Shock Heavy Shock 0.50
0.80
1.25
0.80
1.00
1.50
1.00
1.25
1.75
1.25
1.50
2.00
0.80
1.00
1.50
1.00
1.25
1.75
1.25
1.50
2.00
1.50
1.75
2.25
1.00
1.25
1.75
1.25
1.50
2.00
1.50
1.75
2.25
1.75
2.00
2.50
4.4 when drives are equipped with brakes on the input, and the torque rating of the brake exceeds the rating of the motor, the rating of the brake dictates the selection of the gear unit. 4.5 The recommended service factors for Dry Dock Crane applications are given in Table 3. Due to the nature of these crane drives, the service factors are to be used for any duration of service. 4.6 When a fluid coupling is used between the prime mover and the gear unit, the service factor given in Table 1 for moderate or heavy shock may be modified based on the unit manufacturer's analysis and recommendation for the application. 4.7 The maximum momentary or starting load must not exceed 200 per cent of rated load (1 00 per cent overload). Rated load is defined as the unit rating with a service factor of 1.0. 4.8 The service factors listed for paper mill applications are consistent with those shown in TAPPI (Technical Association of the Pulp and Paper Industry) Standard 406.08, "Service Factors for Gears on Major Equipment in the Paper and Pump Industry." AGMA STANDARD PRACTICE FOR ENCLOSED SPEED REDUCERS OR INCREASERS USING SPUR, HELICAL, HERRINGBONE AND SPIRAL BEVEL GEARS Table 2 Application Classification for Enclosed Speed Reducers Application AGITATORS Pure Liquids Liquids and Solids Liquids-Variable Density
Load Classification
Uniform Moderate Shock Moderate Shock
BLOWERS Centrifugal Lobe Vane
Uniform Moderate Shock Uniform
BREWING AND DISTILLING Bottling Machinery Brew kettles, Cont. Duty
Uniform
Uniform Uniform Cookers-Cont. Duty Mash Tubs-Cont. Duty Scale Hopper, Frequent Starts CAN FILLING MACHINES
Uniform Moderate Shock Uniform
CANE KNIVES
SF = 1.50
CAR DUMPERS
Heavy Shock
CAR PULLERS
Moderate Shock
CLARIFIERS
Uniform
CLASSIFIERS
Moderate Shock
CLAY WORKING MACHINERY Brick Press Briquette Machine Clay Working Machinery Pug Mill
Heavy Shock Heavy Shock Moderate Shock Moderate Shock
COMPRESSORS Centrifugal Lobe Reciprocating, Multi-Cylinder Reciprocating, Single-Cylinder
Uniform Moderate Shock Moderate Shock Heavy Shock
CONVEYORS-UNIFORMALLY LOADED OR FED Apron Assembly Belt Bucket Chain Flight Oven Screw
Uniform Uniform Uniform Uniform Uniform Uniform Uniform Uniform
AGMA STANDARD PRACTICE FOR ENCLOSED SPEED REDUCERS OR INCREASERS USING SPUR, HELICAL, HERRINGBONE AND SPIRAL BEVEL GEARS Table 2 Application Classification for Enclosed Speed Reducers (continued) Application CONVEYORS-HEAVY DUTY NOT UNIFORMLY FED Apron Assembly Belt Bucket Chain Flight *Live Roll Oven Reciprocating Screw
Load Classification
Moderate Shock Moderate Shock Moderate Shock Moderate Shock Moderate Shock Moderate Shock
Moderate Shock Shaker
Heavy Shock Moderate Shock Heavy Shock
CRANES (See Table 3 for Dry Dock Cranes) Main Hoists *Bridge Travel *Trolley Travel
CRUSHER Ore Stone Sugar
Uniform
Heavy Shock Heavy Shock SF = 1.50
Moderate Shock DREDGES Cable Reels Conveyors Cutter Head Drives Jig Drives Maneuvering Winches Pumps Screen Drive Stackers Utility Winches
Moderate Shock Heavy Shock Heavy Shock Moderate Shock Moderate Shock Heavy Shock Moderate Shock Moderate Shock
DRY DOCK CRANES
See Table 3
Uniform ELEVATORS Bucket-Uniform Load Bucket-Heavy Load Bucket-Cont. Centrifugal Discharge Escalators Freight Gravity Discharge *Man Lifts *Passenger
**EXTRUDERS(Plastic) Film Sheet Coating Rods Pipe
Moderate Shock Uniform Uniform Uniform Moderate Shock Uniform
Uniform Uniform Uniform
Uniform Uniform Uniform
Tubing Blow Molders Pre-plasticizers
Moderate Shock Moderate Shock
*Refer to Factory **To be selected on basis of 24hr service only. APPENDIX F. Comparison of capital cost of slipring motor with switchgear versus D.O.L. start S cage motor with scoop coupling and system loses. For 3,3 Kw: 450 Kw
375 Kw
225 Kw
108
118
109
Total Estimated Losses 7%
7%
8%
D.O.L. start
100
100
Slipring
100
Total Estimated Losses 8% to 11% 9% to 12% 10% to 12%
Starting Current
2 * full load
6 * full load
For 500 Volt: Slipring
130
110
103
Total estimated losses 7%
7%
8%
D.O.L. Start
100
100
100
2 * Full load
6 * Full Load
Total estimated losses 9% to 11% 9% to 12% 10% to 12% Please note only one manufacturers prices were used for motors, fluid couplings and switchgear. No allowance is made for supply cable size to cater for starting current.
Screw Conveyor Engineering Guide Introduction The engineering section of this catalog was compiled to aid you in the design of a conveyor system, yielding optimum performance and efficiency, for your individual conveying function. Primary considerations for the selection of a screw conveyor are:
1.
Type and condition of the material to be handled, including maximum particle size, and, if available, the specific bulk density of the material to be conveyed. 2. Quantity of transported material, expressed in pounds or tons per hour. 3. The distance for which the material is to be conveyed.
In the next sections is the necessary information for the selection of a screw conveyor system, presented in a series of five steps. These steps are arranged in logical order, and are divided into separate sections for simplicity. The five steps are: 1. Establishing the characteristics of the material to be conveyed. Locating conveyor capacity (conveyor size and speed) on capacity tables. 3. Selection of conveyor components. 4. Calculation of required horsepower. Checking of component torque capacities (including selection of shaft types and sizes). 2.
5.
All necessary calculations are expressed in graphic and equation form, and use of all charts, graphs, etc. will be explained fully at the end of each section. Engineering data regarding the design of screw feeders and their selection, is presented in a separate section, immediately following the screw conveyor data. Any unusual applications, or special designs, should be referred to KWS Mfg's. Engineering Department.
Screw Conveyor Engineering Guide Material Characteristics The Material Tables on the following pages contain information regarding materials which may be effectively conveyed, using KWS Manufacturing Company's screw conveyor systems. For information on unlisted materials, refer to the Engineering Department of KWS Manufacturing Company, Inc. "Convey-ability" data for unlisted materials can, when necessary, be complied by making a comparison of listed materials which have similar physical characteristics, such as weight and particle size. The following is a brief description of the information presented in the Materials Table.
Maximum Particle Size Conveyor size, speed, and horsepower requirements, are directly affected by bulk density and internal friction, which are relative to the particle size of conveyed material,
Average Weight per Cubic Foot This section of the Materials Table is supplied to enable you to convert the required capacity in pounds or tons per hour to volume in cubic feet per hour. Note: Since most typical applications of screw conveyors receive slightly aerated, gravity fed products, the weights listed in this table are averages and when possible, actual bulk densities should be used.
Conveyor Loading
The recommended percentages of conveyor loading is a prime factor in determining the size of conveyor, and is based on the maximum depth at which materials will flow through the conveyor without causing undue wear. Considerations should be made, for example, for materials with a high abrasiveness because wear indexes would normally be higher due to a larger contact area with component parts.
Horsepower Factor The horsepower factor, representing the relative mobility of the material, is necessary for horsepower calculation.
Recommended Component Series This information is presented to assist in the selection of the proper materials of construction, component weights and other specifications best suited for the material to be conveyed. The alphabetical code refers to general component series, and the numerical code gives bearing and shaft type recommendations. See component series tables
Abrasiveness, Corrosiveness And "Flow-ability" In addition to the above information, the Materials Table also presents graphically the relative abrasiveness, corrosiveness and "flow-ability" of the materials listed. These characteristics, as well as other special aspects of materials, are given further treatment in the Component Selection Section. The values of the graphic presentation used in the Materials Table are listed below. Description
I
ll
III
Abrasiveness
Not Abrasive
Mildly Abrasive
Highly Abrasive
Corrosiveness
Not Corrosive
Mildly Corrosive
Highly Corrosive
Free Flowing
Relatively Free Flowing
Sluggish
Flowability
Angle of Repose To 30°
30° - 45°
Beyond 45°
Note: Some materials, while they are not corrosive under "normal" conditions; may become corrosive under certain other conditions, such as when heated or in the presence of moisture.
Special Characteristics Notes Notable unusual material characteristics are fisted by numerical codes in the last column of the table where applicable. An explanation of these numerical codes is given below. 1.
Contains explosive dust 2. Fluidizes easily 3. Absorbs moisture. 4. Usage or value affected by contamination. 5. Emits toxic fumes or dust. 6. Usage or value affected by material degradation. 7. Exceptionally light or fluffy. 8. Tends to pack under pressure. 9. Fibrous material which tends to mat.
Materials Table
A | B | C | D-G | H-L | M-O | P-R | S | T-Z Tables are in alphabetical order. Click above to navigate to different sections. Table keys and explanations are located in the "Material Characteristics" Section.
Material
Maximum Average H.P. Component Abras- Corrosi- FlowParticle Size Weight Per Loading Note Factor Series iveness veness ability (IN.) Cu.Ft.
Acetylenogen (Calcium Carbide)
+ 1/2
70-80
30B
1.6
B4
II
I
II
1
Adipic Acid
-100M
45
30A
0.8
D3
I
II
II
3
Alfalfa Meal
-1/8
17
30A
0.9
B4
II
I
III
.7
Alfalfa Seed
-1/8
48
30B
0.5
B4
II
I
I
1
Almonds
-1/2
28-30
30B
0.9
B4
II
I
II
6
Alum
-1/8
45-58
30A
0.6
A2
I
I
II
•
Alum, lumpy
+ 1/2
50-60
30A
1.4
B1
I
I
II
•
Alumina
-100M
60-120
15
1.8
C4
III
I
I
2
Aluminate Gell, dried
-100M
45
30B
1.7
B4
II
I
II
Aluminum Chips
-1/2
7-15
30A
0.8
A2
I
I
III
Aluminum Hydrate (Aluminum Hydroxide)
-1/2
13-18
30A
1.4
A2
I
I
III
Aluminum Oxide (Alumina)
-100M
60-120
15
1.8
C4
III
I
I
Aluminum Ore (Bauxite)
-3
75-85
15
1.8
D4
III
I
II
Aluminum Silicate
-1/8
49
45
0.8
A2
I
I
II
Aluminum Sulfate (Alum)
•
•
•
•
•
•
•
•
Amianthus (Asbestos )
Fibers
20-40
30B
1.0
B4
II
I
III
Ammonium Chloride, Crystalline
-1/8
52
30A
0.8
A2
I
I
II
Ammonium Nitrate
-1/8
45-62
•
•
•
•
•
•
• 1.3
Ammonium Sulfate
•
40-58
•
•
•
•
•
III
•
Andalusite (Aluminum Silicate)
-1/8
49
45
0.8
A2
I
I
II
9
2
5,7,8
Antimony
-100M
•
30B
•
B4
II
I
II
•
Apple Pomace, dry
-1/2
15
30B
0.5
B4
II
I
III
7
Arsenate of Lead (Lead Arsenate)
-1/8
72
30A
1.0
A2
I
I
III
2, 5
Arsenic
-100M
30
•
•
•
•
•
•
•5
Arsenic Oxide (Arsenolite)
•
100-120
•
•
•
•
•
•
•5
Asbestos, Ore
-1/2
81
15
1.2
C4
III
I
II
5
Asbestos, Shred
Fibers
20-40
30B
1.0
B4
II
I
III
5, 7, 8
Ashes, Coal, dry
-1/2
35-45
30B
2.0
B4
II
I
III
Ashes, Coal, dry
-3
35-45
30B
2.0
B4
II
I
III
Ashes, Coal, wet
-1/2
45-50
30B
3.0
D4
II
II
III
8
Ashes, Coal, wet
-3
40-50
15
4.0
D4
II
II
III
8
Asphalt, Crushed
-1/2
45
30A
2.0
A2
I
I
II
Conveyor Capacity A capacity table is provided on the next section to aid you in calculation of proper conveyor size. To use this table, find the capacity at maximum RPM, opposite the recommended percentage of conveyor loading, which equals or exceeds the capacity of material required per hour. The recommended conveyor diameter will then be found in the appropriate column on the same line, as will the maximum particle size recommended for the screw diameter. If the maximum particle size you plan to convey is larger than the maximum recommended particle size for the conveyor you've chosen from the table, you must then select a larger conveyor, adequate to handle the maximum particle size you intend to use.
Calculation of Conveyor Speed Conveyor speed can be most conveniently calculated, by use of the nomographs supplied on pages To use this nomograph first locate the two known values (screw diameter, and required capacity, in cu. ft. per hour) then with a straight edge connect these two points, and the appropriate conveyor speed will be the intersection point on the third value column marked "speed". Maximum economical capacities will be listed for reference opposite their respective conveyor diameters, and should not be exceeded. Another method of calculating conveyor speed is: CS =
CFH CFH at 1RPM
Equation Symbols CS = Conveyor Speed CFH = Capacity in Cubic Feet per Hour
Capacity Factors for Special Pitch or Modified Flight Conveyors Special conveyor types are selected in the same manner as standard conveyors, but the section capacity used for determining size and speed, must be modified to compensate for different characteristics of special conveyors. Calculation of special screw conveyor capacities is as follows: SC = CFH x CF
Equation Symbols
SC = Selection Capacity CFH = Required Capacity in Cubic Feet per Hour CF = Capacity Factor
Special Conveyor Pitch Capacity Factors Pitch
Description
Capacity Factor
Standard
Pitch = Diameter
1.00
Short
Pitch = 2/3 Diameter
1.50
Half
Pitch = 1/2 Diameter
2.00
Long
Pitch = 1-1/2 Diameters
0.67
Special Conveyor Flight Capacity Factors Conveyor Loading
Type
15%
30%
45%
95%
Cut flight
1.92
1.57
1.43
*
Cut & folded flight
*
3.75
2.54
*
Not Recommended Factors for Conveyors With Paddles* Paddles Per Pitch Factor
1
2
3
4
1.08
1.16
1.24
1.32
* Std. paddles at 45° reverse pitch
Ribbon Conveyor Capacity Factors Conveyor Loading
Dia.
Ribbon Width
15%
30%
45%
6
1
1.32
1.52
1.79
9
1-1/2
1.34
1.54
1.81
10
1-1/2
1.45
1.67
1.96
12
2
1.32
1.52
1.79
2-1/2
1.11
1.27
1.50
14
2-1/2
1.27
1.45
1.71
16
2-1/2
1.55
1.69
1.90
18
3
1.33
1.53
1.80
20
3
1.60
1.75
1.96
24
3
2.02
2.14
2.28
Example:
A conveyor is required to transport 10 tons per hour of a material weighing 62 pounds per cubic foot and having a maximum particle size of 100 mesh. To further complicate the problem, we will require that the material be mixed in transit using cut and folded flights. Since the distance the material is to be conveyed is relatively short, we want to use short pitch screws, to insure proper mixing of material. The materials table recommends a loading percentage of 30% A.
Actual calculated volume: 20,000 lbs. 62 lbs. / cu. ft.
= 323 cu. ft./hr.
For proper calculation of size and speed, this volume must be corrected, by use of capacity factors, to compensate for cut and folded, and short (2/3) pitch flights. These capacity factors, taken from the preceding charts are: Cut and folded flights 30% loading = 3.75 Short pitch flights (2/3 pitch) = 1.50 With capacity factors included, capacity will now be calculated: SC = 3.75 * 1.50 * 323 SC = 1817cu.ft. This selection capacity value will be used in the capacity table, for calculating correct size and speed. In the appropriate column, under 30% A loading, we find that a 14" conveyor, at the maximum recommended speed will convey 2194 cu. ft. per hr. or 21.1 cu. ft. per revolution. To calculate actual conveyor speed, the following formula should be used: 1817 cu. ft. / hr. 21.1 cu. ft. / hr. at 1 RPM
= 86.1 RPM
This is the correct speed at which the 14" conveyor with cut and folded, and short pitch flights will convey the actual capacity of 323 cu. ft. per hour. Graphic selection of this conveyor could also be accomplished by use of the 30% A nomograph on page 22 and the selection capacity of 1817 cu. ft. per hour.
Conveyor Capacity Table Trough Loading
Screw Dia.
Max. Lump Size (In.)
Max. RPM
4
1/2
6
3/4
9
Capacity in Cu. Ft. Per Hr. At Max. RPM
At 1 RPM
69
14.5
.21
66
49.5
.75
1-1/2
62
173
2.8
10
1-3/4
60
222
3.7
12
2
58
389
6.7
14
2-1/2
56
588
10.5
16
3
53
832
15.7
18
3-1/4
50
1,135
22.7
20
3-1/2
47
1,462
31.1
24
4
42
2,293
54.6
4
1/2
139
57
.41
6
3/4
132
198
1.5
9
1-1/2
122
683
5.6
10
1-3/4
118
849
7.2
12
2
111
1,476
13.3
14
2-1/2
104
2,194
21.1
16
3
97
3,046
31.4
18
3-1/4
90
4,086
45.4
20
3-1/2
82
5,092
62.1
24
4
68
7,426
109.2
4
1/2
69
28
.41
6
3/4
66
99
1.5
9
1-1/2
62
347
5.6
10
1-3/4
60
432
7.2
12
2
58
771
13.3
14
2-1/2
56
1,182
21.1
16
3
53
1,664
31.4
18
3-1/4
50
2,270
45.4
20
3-1/2
47
2,919
62.1
24
4
42
4,586
109.2
4
1/2
190
116
.61
6
3/4
182
413
2.27
9
1-1/2
170
1,360
8.0
10
1-3/4
165
1,782
10.8
12
2
157
3,030
19.3
14
2-1/2
148
4,558
30.8
16
3
140
6,524
46.6
18
3-1/4
131
8,659
66.1
20
3-1/2
122
11,590
95.0
24
4
105
17,535
167.0
95% Loaded Conveyors
Conveyor loadings may sometime exceed the recommended % of Loading, listed in the materials table. Considerations as to the material characteristics may justify up to 95% loading of tubular or shrouded conveyors. The following table lists maximum speeds limited with regard to the percentage of loading normally recommended for the specific listed materials.
Capacity Table For 95% Loaded Conveyors Max. Recommended RPM Screw Dia. Max. Lump Size (IN.)
Normal % Loading * 15
30A
30B
45
Capacity in Cubic Feet Per Hour Normal % Loading * 15
30A
30B
45
AT 1 RPM
4
1/4
76
89
80
96
96
113
101
122
1.27
6
3/8
67
78
70
84
318
370
332
399
4.75
9
3/4
58
68
61
73
974 1,142 1,024 1,226
16.8
10
7/8
55
65
58
70
1,309 1,547 1,380 1,666
23.8
12
1
49
58
52
62
1,999 2,366 2,122 2,530
40.8
14
1-1/4
43
51
46
55
2,804 3,325 2,999 3,586
65.2
16
1-1/2
38
45
40
48
3,762 4,455 3,960 4,752
99.0
18
1-3/4
32
38
34
41
4,512 5,358 4,794 5,781
141.0
20
2
26
31
28
34
5,226 6,231 5,628 6,834
201.0
24
3
21
25
23
28
7,434 8,850 8,142 9,912
354.0
Basic Conveyor Flight & Pitch Types
Standard Pitch, Single Flight Conveyor screws with pitch equal to screw diameter are considered standard. They are suitable for a wide range of materials in most conventional applications.
Short Pitch, Single Flight Flight pitch is reduced to 2/3 diameter. Recommended for inclined or vertical applications. Used in screw feeders. Shorter pitch retards flushing of materials which fluidize.
Half Pitch, Single Flight Similar to short pitch, except pitch is reduced to 1/2 standard pitch. Useful for vertical or inclined applications, for screw feeders and for handling extremely fluid materials.
Long Pitch, Single Flight Pitch is equal to l½ diameters. Useful for agitating fluid materials or for rapid movement of very free-flowing materials.
Variable Pitch, Single Flight Flights have increasing pitch and are used in screw feeders to provide uniform withdrawal of fine, freeflowing materials over the full length of the inlet opening.
Double Flight, Standard Pitch Double flight, standard pitch screws provide smooth, regular material flow and uniform movement of certain types of materials.
Tapered, Standard Pitch, Single Flight Screw flights increase from 2/3 to full diameter. Used in screw feeders to provide uniform withdrawal of lumpy materials. Generally equivalent to and more economical than variable pitch.
Single Cut-Flight, Standard Pitch Screws are notched at regular intervals at outer edge. Affords mixing action and agitation of material in transit. Useful for moving materials which tend to pack.
Cut & Folded Flight, Standard Pitch Folded flight segments lift and spill the material. Partially retarded flow provides thorough mixing action. Excellent for heating, cooling or aerating light substances.
Single Flight Ribbon Excellent for conveying sticky or viscous materials. Open space between flighting and pipe eliminates collection and buildup of the material.
Standard Pitch With Paddles Adjustable paddles positioned between screw flights oppose flow to provide gentle but thorough mixing action.
Paddle Adjustable paddles provide complete mixing action, and controlled material flow.
Conveyor Component Selection Proper selection of components is very important in the design of conveyor system. This section of the Engineering Catalogue explains the different designs of primary components, and their principle uses. Also, there is a list of special influencing factors for materials with special handling characteristics.
Conveyor Loading and Discharge Conveyor loading should be regulated to prevent the components from exceeding their design limits.
Regulated Output Devices When delivery to conveyor is from machinery with a regulated material output, the conveyor itself can be designed to handle the anticipated material volume. Material is sometimes stored and released intermittently. In this situation, surge loads sometimes cause the conveyor to operate beyond its recommended capacity. Screw feeders are very effective in regulating these intermittent loads, and should be used if at all possible. Otherwise conveyors must be designed for the maximum momentary or surge loads.
Static Storage Loading When loading from static storage or from manually regulated inlets, a load indicating ammeter can be attached to the meter control, as a simple and effective tool for accomplishing maximum design loading.
Multiple Inlet Loading When more than one inlet is feeding the screw conveyor, care must be taken to insure the collective total of the inlets does not exceed the conveyors design limits.
Loading Through Automatic Controls Automatic devices are available to modulate inlet or feed devices to work within design limits of the conveyor at all times.
Discharge Methods Below are drawings of standard discharge components in a variety of designs. These configurations are listed for individual applications where the standard discharge spouts are not necessarily appropriate. Cautions are inserted when necessary for particular discharge components.
Standard Discharge Spout This component provides a means of directly coupling most interconnecting spouts, processing machinery, other conveyors or storage facilities. Available with hand, rack and pinion or air actuated cut off gates.
Flush End Discharge
Mechanically Operated Gates (Manual or Remote Controlled) Caution: Not to be used on last discharge.
Plain Opening
Open Bottom Discharge
Trough End Discharge Caution: To be used with conveyor loadings of 30% or under only.
Open End Discharge
Factors Influencing Materials of Construction or Special Mechanical Arrangement of Screw Conveyor Components
Corrosive Materials Corrosive materials, or materials which have a tendency to become corrosive under certain conditions, may necessitate the use of corrosion-resistant alloys such as stainless steel.
Abrasive Materials Abrasive materials which may cause excessive wear of components should be conveyed at a nominal depth in the conveyor. It is often advisable to also specify KWS's Abrasion-Resistant Screw Conveyors or conveyors with flights formed of AR steel plate. KWS's Abrasion-Resistant Screw Conveyors, which have a Rockwell C hardness of 68-70, are covered in the Component Section. A table listing the standard width of application of hard surfacing is included.
Contaminable Materials Materials whose usefulness or value may be altered by contamination may require the use of non-lubricated bearings, as well as a tightly sealed system.
Hygroscopic Materials Materials that readily absorb moisture require a tightly sealed conveyor. It may be necessary also to jacket the conveyor trough or housing with a circulating medium to maintain an elevated temperature. Purging the system with dry gas or air may be necessary.
Interlocking Materials Materials which tend to mat or interlock are sometimes effectively conveyed by using special devices to load the conveyor.
Fluidizing Materials Some materials tend to assume hydraulic properties when aerated or mechanically agitated. Such materials may "flow" in the conveyor much the same as a liquid. These materials should be referred to KWS's engineering department for recommendations.
Explosive Materials Dangerous explosive materials can be handled by sealing the system and/or the use of non-sparking components. It is also possible to utilize exhaust systems for hazardous dust removal.
Materials Which Tend to Pack Materials that tend to pack under pressure can frequently be handled by using aerating devices (for fine materials) or special feeding devices (for large or fibrous particles).
Viscous or Sticky Materials Viscous or sticky materials are transported most effectively by ribbon conveyors.
Degradable Materials Materials with particles that are easily broken may be effectively handled by selection of a larger, slower conveyor.
Elevated Temperatures Materials handled at elevated temperatures may require components manufactured of high-temperature alloys. If it is feasible to cool the material in transit, a jacketed trough used as a cooling device may also be employed.
Toxic Materials Materials which emit harmful vapors or dusts require tightly sealed systems. Exhaust devices may also be used to remove the vapors or dusts from the conveyor housing.
Description of Components Conveyor Screws The recommended screws listed in the Component Series Table are standard KWS helicoid and sectional screw conveyors. The use of helicoid or sectional conveyors is largely a matter of individual preference. It is advisable to use, whenever possible, standard conveyors in standard lengths. When a special short length must be used to make up the total conveyor length, it is preferably located at the discharge end.
Screw conveyors are structurally reinforced at the ends by the use of end lugs which are welded to the noncarrying side of the flights so that material flow will not be obstructed. Screw conveyors which move material in a single direction should not be turned end-for-end unless the direction of screw rotation is reversed. Likewise, the direction of rotation should not be reversed unless the conveyor is turned end-for-end. Requirements for reversible conveyors should be referred to KWS's Engineering Department. Flighting should be omitted at the final discharge, so that material will not carry past the discharge point.
To assure proper material flow past hanger bearing points, flight ends should be positioned to each other at 180 degrees. The "hand" of a conveyor, in conjunction with the direction the conveyor is rotated, determines the direction of material flow. The diagram below illustrates flow direction for "right-hand" and "left-hand" conveyors when rotated clockwise or counterclockwise. A right-hand screw conveyor pulls the material toward the end which is being rotated clockwise. The direction of flow is reversed when the direction of rotation is reversed. A left-hand conveyor pushes the material away from the end which is being rotated clockwise. Again, the direction of material flow is reversed when the direction of rotation is reversed. To determine the hand of a conveyor, observe the slope of the near side of the flighting. If the slope is downward to the right, the conveyor is right-hand. If the slope is downward to the left, the conveyor is lefthand. Right-hand conveyor is furnished unless otherwise specified.
Troughs and Tubular Housings KWS troughs and tubular housings are available in standard lengths. Special lengths are available when required. All conveyor troughs or tubular housings should be supported by flange feet or saddles at standard intervals. Extreme end flanges should be supported with feet so that the conveyor ends may be removed without disturbing trough or housing alignment.
Inlets and Discharges The proper methods of conveyor loading and discharge were covered previously in this section.
Shafts The primary consideration in determining the type and size of coupling and drive shafts is whether the shafts selected are adequate to transmit the horsepower required, including any overload. Normally, cold-rolled shafts are adequate. However, high-tensile shafts may be required due to torque limitations. Also, stainless steel shafts may be necessary when corrosive or contaminable materials are to be handled. Conveyors equipped with non-lubricated iron hanger bearings require hardened coupling shafts, and hard-surfaced hanger bearings require hard-surfaced shafts. Specific shaft size determination is covered in the Torque Capacities Section.
Shaft Seals Several conveyor end seal types are available to prevent contamination of the conveyed material or to prevent the escaped of material from the system.
Bearings Hanger Bearings - The purpose of hanger bearings is to provide intermediate support when multiple screw sections are used. Hanger bearings are designed primarily for radial loads. Adequate clearance should be allowed between the bearings and the conveyor pipe ends to prevent damage by the thrust load which is transmitted through the conveyor pipe. The hanger bearing recommendations listed in the Component Series Table are generally adequate for the material to be handled. Often, however, unusual characteristics of the material or the conditions under which the conveyor must operate make it desirable to use special bearing materials. A list of available special bearing materials is provided in this section. For specific recommendations regarding the use of special bearing materials, consult KWS's Engineering Department.
End Bearings - Several end bearing types are available, and their selection depends on two basic factors: Radial load and thrust load. The relative values of these loads determines end bearing types.
Radial load is negligible at the conveyor tail shaft. However, drive ends (unless integrated with the conveyor end plate) are subject to radial loading due to overhung drive loads, such as chain sprockets or shaftmounted speed reducers. Thrust is the reaction, through the conveyor screw or screws, resulting from movement of the material. Therefore, the end bearing must prevent axial movement of the screw which would allow contact with hanger bearings or ends. Thrust bearings should be located at the discharge end of the conveyor. This places the conveyor in tension, preventing deflection in the screws when the system is heavily loaded. The following diagrams illustrate discharge and inlet end positions of the thrust bearing.
Component Series The recommended Component Series for the material to be conveyed may be found in the Materials Table at the beginning of the Engineering Section. The alphabetical code relates to the general component series, and the numerical code refers to bearings and coupling shafts. The Component Series Table follows on the next page. Bearing and coupling shaft recommendations are listed in the table below. The Component Series Table lists the screw conveyor numbers for both helicoid and sectional screws and gives the trough and cover thicknesses. The Bearing and Coupling Shaft Table lists the recommended materials of construction. Series
Coupling Shaft
Bearing Material
1
Standard or High Torque
Babbitt Wood Bronze
2
Standard or High Torque
Babbitt Wood Bronze Ball
3
Standard or High Torque
Babbitt
4
Hardened or Hard Surfaced
Hard Iron Hard Surfaced
Conveyor Component Series Table Other Bearing Materials Available
•
•
• Graphite Bronze Graphite-Impregnated Plastic • Machined Nylon • Molded Nylon • Oil-Impregnated Bronze Plastic, Laminated Fabric-Base • Teflon
Component Series Table Screw Shaft Cover Dia. Dia. Thickness
Series A Screw Number Helicoid
Sectional
Series B
Tube or Trough Thickness
Screw Number Helicoid
Sectional
Tube or Trough Thickness
4
1
16 Ga.
4H204
16 Ga.
4H206
6
1 1/2
16
6H304
6S309
16
6H308
6S309
14
9
1 1/2 2
16
9H306 9H406
9S309 9S409
14
9H312 9H412
9S309 9S409
10
10
1 1/2 2
16
10H306 10H412
10S309 10S409
14
10H306* 10H412
105312 10S412
10
12
2 2 7/16 3
14
12H408 12H508 12H614
12S409 12S509 12S609
12
12H412 12H512 12H614
12S412 12S512 12S612
3/16"
14
2 7/16 3
14
14H508 14H614
14S509 14S609
12
14H508 14H614
14S512 14S612
3/16"
16
3
14
16H610
16S612
12
16H614
16S616
3/16"
18
3 3 7/16
14
18S612 18S712
12
18S616 18S716
3/16"
20
3 3 7/16
14
20S612 20S712
10
20S616 20S716
3/16"
24
3 7/16
12
24S712
10
24S716
3/16"
Screw Shaft Cover Dia. Dia. Thickness
Series C Screw Number Helicoid
Sectional
14 Ga.
Series D
Tube or Trough Thickness
Screw Number Helicoid
14 Ga.
4H206*
Sectional
Tube or Trough Thickness
4
1
16 Ga.
4H206
6
1 1/2
16
6H312
6S312
14
6H312
6S316
10
9
1 1/2 2
16
9H312 9H414
9S312 9S412
10
9H312 9H414
9S316 9S416
3/16"
10
1 1/2 2
16
10H306* 10H412
10S312 10S412
10
10H306* 10H412*
10S316 10S416
3/16"
12
2 2 7/16 3
14
12H412 12H512 12H614
12S416 12S516 12S616
3/16"
12H412* 12H512* 12H614
12S424 12S524 12S624
1/4"
14
2 7/16 3
14
14H508* 14H614
14S524 14S624
3/16"
14H508* 14H614*
14S524 14S624
1/4"
16
3
14
16H614
16S616
3/16"
16H614*
16S624
1/4"
18
3 7/16
14
18S624 18S724
3/16"
18S624 18S724
1/4"
20
3 3 7/16
14
20S624 20S724
3/16"
20S624 20S724
1/4"
24
3 7/16
12
24S724
3/16"
24S724
1/4"
*Hard-Surfacing Recommended
Horsepower Calculation
Graphic Method of Calculation The total horsepower (TSHP) required at the drive shaft to drive the loaded conveyor system may be calculated graphically by use of the nomographs at the end of this section. The friction horsepower (FHP), determined with the first nomograph, added to the Material Horsepower (MHP), determined with the second nomograph, equals the Total Shaft Horsepower (TSHP). Friction Horsepower - A straight edge placed at the first two known values, conveyor size (related to hanger bearing class as listed in hanger bearing factor table) and length, will intersect a reference point on the centerline. A straight edge placed from this reference point to the third known value, conveyor speed, will intersect the unknown value, Friction Horsepower, on the last line. Material Horsepower - A straight edge placed at the first two known values, conveyor capacity and Material Horsepower Factor, will intersect a reference point on the centerline. A straight edge from the reference point to the third known value, conveyor length, will intersect the unknown value, Material Horsepower, on the last line.
Calculation by Equation TSHP may also be calculated by equation using the following formulas:
Friction H.P. Calculation: FHP =
DF x HBF x L x S 1,000,000
Material H.P. Calculation: MHP =
CFH x W x MF x L 1,000,000
OR MHP =
CP x MF x L 1,000,000
Note: If calculated Material Horsepower is less than 5 it should be corrected for potential overload. The corrected horsepower value corresponding to the calculated Material Horsepower will be found on the lower scale of the Material. Horsepower Overload Correction Chart.
Total Shaft H.P. Calculation
TSHP = FHP + MHP* *Corrected if below 5 HP. Note: The actual motor horsepower required to drive the loaded conveyor system is dependent on the method used to reduce the speed the motor to the required speed of the conveyor. Drive losses must be taken into consideration when selecting the motor and drive equipment.
Equation Symbols TSHP
Total Shaft H.P.
FHP
Friction H.P. (H.P. required to drive the conveyor empty)
MHP
Material H.P. (H.P. required to move the material)
L
Conveyor Length
S
Conveyor Speed
DF
Conveyor Diameter Factor
HBF
Hanger Bearing Factor
CFH
Conveyor Capacity
W
Weight per cu. ft.
CP
Capacity, lbs. per hr.
MF
Material H.P Factor (From the Materials Table)
Diameter Factors Diameter
Factor
4
12
6
18
9
31
10
37
12
55
14
78
16
106
18
135
20
165
24
235
Hanger Bearing Factors Bearing Type
Bearing Factor
Bearing Class
Ball
1.0
I
Babbit
1.7
II
2.0
III
Bronze *Graphite Bronze Plastic, laminated fabric-base Nylon * Bronze, oil-impregnated Wood *Plastic, graphite- impregnated
*Nylon *Teflon *Hard Iron
4.4
IV
*Hard-Surfaced
*Non-Lubricated
Conveyors With Modified Flights The procedure for calculation of horsepower for conveyors with special or modified flights is identical to that used for standard conveyors except that the Material Horsepower must be multiplied by one or more of the following applicable factors.
Modified Flight Factors Conveyor Loading
Flight Type
15
30
45
95
1.10
1.15
1.2
*
Cut & Folded Flight
*
1.50
1.7
*
Ribbon Flight
1.05
1.14
1.20
*
Cut Flight
*Not Recommended Conveyors With Paddles*
Conveyors With Paddles Paddles Per Pitch Factor
1
2
3
4
1.29
1.58
1.87
2.16
* Std. paddles at 45° reverse pitch Total Shaft Horsepower (TSHP) is calculated by adding Material Horsepower, multiplied by the appropriate modified flight factor or factors, to Friction Horsepower. Note: Conveyors which have deviation in pitch only do not require special consideration, and their horsepower calculations are as described for standard conveyors.
Example A 10-inch conveyor 35 feet long with a capacity of 10 tons per hour at 45 RPM has been selected. From the Materials Table, a Horsepower Factor of 0.8 is found for the material to be conveyed. The table also indicates Series 4 hanger bearings and shafts. Hard iron bearings and hardened coupling shafts have been selected to suit this requirement. Friction Horsepower, the horsepower required to drive the conveyor empty, is calculated as follows: Diameter Factor = 37 Hanger Bearing Factor = 4.4
Length = 35 RPM = 45 FHP =
37 x 4.4 x 35 x 45 1,000,000
= 0.256
Material Horsepower, the horsepower required to move the material, is calculated by the following equation: Capacity (in lbs. per hr.) = 20,000 Horsepower Factor Length RPM = 4.4 Length = 35 RPM = 45 MHP =
20,000 x 0.8 x 35 1,000,000
= 0.560
Since the calculated Material Horsepower is less than 5, it is necessary to find the corrected horsepower value corresponding to 0.56 horsepower on the Overload Correction Chart below. This value is found to be 1.320 horsepower. Total Shaft Horsepower (TSHP) is the sum of Friction horsepower and the corrected Material Horsepower. Thus TSHP is calculated as follows: TSHP = 0.256 + 1.320 = 1.576 H.P. Assuming a drive efficiency of 85% resulting in a total drive horsepower of 1.853, a standard 2 horsepower motor would be selected for the drive input. The horsepower required for the above conveyor may also be determined graphically by the use of the two horsepower nomographs. The first nomograph determines Friction Horsepower. The second determines Material Horsepower. Total Shaft Horsepower is determined by adding the two values.
Friction Horsepower
Material Horsepower
Corrected Material H.P.
Friction Horsepower Nomograph
Bearing Class / Conveyor Size
Length
Speed
Friction H.P.
Material Horsepower Nomograph
Capacity Per Hour
Material H.P. Factor
Length
Material H.P.
Conveyor Torque Capacities Although a given conveyor may be adequate insofar as material conveying capacity is concerned, the horsepower available to operate the system may exceed the torque capacities of standard components during overloaded or stalled conditions. To insure adequate torque capacities without undue additional cost, means are provided in the Industrial standard series of conveyor components for more than one maximum allowable horsepower value. This is accomplished by not only a choice of power-transmitting component sizes but also of the materials of construction.
Analysis of a specific conveyor system with regard to component torque adequacy may be conveniently and quickly made by use of the two following nomographs.
Carbon Steel Conveyors The first nomograph covers carbon and high-tensile steel coupling bolts and shafts (drive and coupling) and conveyor pipe (in Schedule 40 and for high capacity, Schedule 80). These components are listed according to their associated standard conveyor shaft diameters. The following table lists actual nominal pipe diameters corresponding to the standard conveyor shaft diameter. Shaft Diameter
1
1-1/2
2
2-7/16
3
3-7/16
Nominal Pipe Size
1-1/4
2
2-1/2
3
3-1/2
4
Stainless Steel Conveyors The second nomograph covers stainless steel coupling bolts, shafts and conveyor pipe. Coupling bolts are listed by the corresponding standard conveyor shaft diameter with which they are used. Conveyor pipes are listed in both Schedule 40s and 80s by their nominal pipe sizes. The pipe size selected should correspond to the standards listed for carbon steel pipe. Deviations from this standard are sometimes possible, in sectional conveyors, by the use of smaller pipe sizes (for economy) when the torque rating is adequate. This procedure requires reaming of the pipe bore for shaft insertion rather than the use of a bushing. It is recommended that requirements for such conveyors be referred to KWS's Engineering Department.
Note:High starting torque motors must not be used without design verification by KWS's Engineering Department. Example: A 12-inch carbon steel conveyor has been selected with a required shaft horsepower of 8.9 and a speed of 64 RPM. The drive to be used has an efficiency of 85%, thus requiring a drive input of 10.46 horsepower. Therefore, a 15 horsepower motor must be used. This total motor horsepower could be transmitted to the conveyor components if overloaded or stalled.
Three standard shaft sizes are available for 12-inch conveyors. They are 2", 2-7/16" and 3". A straight edge is placed from 15 horsepower on the left scale to 64 RPM on the center scale. Project the straight line to the left vertical line of the chart at the right. A horizontal line from this point will pass through component groups suitable for the torque. For the conveyor under consideration, it is found that standard components will be adequate, with the exception of coupling bolts which must be high tensile.
Torque Capacities for Carbon Steel
Drive Motor H.P.
Speed
Component Size*
*Listed sizes based on conveyor shaft diameter
S = standard carbon steel H = high-tensile steel 40 = pipe schedule (standard) 80 = pipe schedule (high capacity)
Torque Capacities for Stainless Steel
Drive Motor H.P.
Speed
Component Size*
*Coupling bolt sizes based on conveyor shaft diameters. Conveyor pipe listed as nominal pipe size.
Conveyor Drive Arrangements KWS offers a complete line of power-transmission equipment. Local distributors provide us with a large stock inventory.
Numerous combinations and types of drives are available for screw conveyor equipment. Some of the more frequently used drives and mechanical arrangements are described below:
Screw Conveyor Drives
A screw conveyor drive consists of a standard single or double reduction shaft-mounted speed reducer, a steel motor mounting bracket, an adapter with CEMA drilling containing shaft seals, and a removable steel shaft, all mounted on a screw conveyor trough end. The motor bracket is rigidly mounted with clearance over the trough end for easy trough cover removal without disassembling any part of the drive. A variety of mounting arrangements makes it possible to locate the drive to avoid interference with other equipment. Correct V-belt tension can be easily maintained by simple adjustment of the motor mounting plate. The drive assembly can be quickly removed by removing the bracket mounting bolts.
Shaft Mounted Drives
The helical gear shaft mount speed reducer uses the screw conveyor drive shaft as an "output shaft," making a mounting base and low speed coupling unnecessary. Because it does not mount to the trough end it offers several advantages. It can be used in limited, higher temperature applications where damaging heat can be dissipated before it affects the reducer. You have a greater variety of seals and bearings to choose from. You can utilize heavy duty bearing for higher than usual bearing loads. The reducer can be rotated in any position around the shaft.
V-belt tension is maintained in the same manner as the screw conveyor drive when using the adjustable motor mount. A tie-rod turnbuckle locks the shaft-mounted reducer into position. (We believe this is best accomplished in the field, consequently we do not normally support the tie-rod from the conveyor.)
Combination Motor-Reducer
Integral motor-reducer drives consist of a combination motor and speed reducer which may be mounted directly to the conveyor cover with an adapter base. The motor-reducer may also be mounted in other positions, depending on available space and accessibility. The motor-reducer output shaft is connected to the conveyor drive shaft through roller chain and sprockets. Speed changes in the field are possible by replacement of one or both sprockets. Suitable conveyor drive end bearings are required for the overhung sprocket loads.
Other Drives Other drive equipment which may be required includes variable speed units which allow manual or automatic adjustment of conveyor capacity by speed deviation. Such drives are especially useful for regulating the flow of material into a process.
Note Fluid, pneumatic or resilient couplings may be used for starting heavily loaded conveyors and to prevent drive component damage due to heavy intermittent overloads.
KWS Belt Guards
Our belt guards are custom designed to meet your specific requirements. They are O.S.H.A. approved and will accent the best of drives. Our two piece construction provides you with the best available features. The back panel is designed to be securely supported. The front panel with sides is easily removable by loosening a hand knob. This permits complete access to sheaves, bushing and V-belts. Standard Features:
•
Painted O.S.H.A. yellow enamel
• Slotted for belt adjustment • 16 ga. steel construction • Fully enclosed • Safe - rounded ends R = Driver, ½" P.D. +2"
CD = Center Distance
SH = Shaft Diameters ½"
R1 = Driver, ½" P.D. +2"
W = Longest Hub+2" (min 4")
SH1 = Shaft Diameter+ 1"
Screw Feeders
1.
3.
Inlet opening matches bin or hopper discharge. 2. Shroud cover prevents material flooding. Twin tapered, variable pitch Screw Conveyor permits even draw off of material. 4. Twin tapered trough. Also available with drop bottom feature. 5. Discharge opening. 6. Solid shafting transmits rotary motion to driving gears. 7. Driving gears synchronize the action of the screw conveyors.
Normally short in length, Screw Feeders are designed to regulate the volumetric rate of material flow from a hopper, bin or storage unit. The inlet is usually flooded with material (100% load capacity) but by incorporating changes in the construction of the flighting (diameter, pitch, etc.) and the speed of the feeder screw, the material discharge can be governed to the desired rate. Feeders can be built with variable diameter or stepped pitch or both in units (com-posed of one, two or a multiple number of screws (i.e., Live Bottom Bin) depending on the application. Screw Feeders are normally equipped with a shroud (curved) cover for a short distance beyond the inlet opening. This prevents flooding of the conveyor with material. When handling very freely flowing materials, extended shroud covers, tubular housing construction or short pitch flights are occasionally required for positive control. Screw Feeders with uniform diameter and pitch normally convey the material from the rear of the inlet opening first. To draw off material evenly across the full length of the inlet, a tapered screw or stepped pitch conveyor screw is required. While Screw Feeders are available in many designs to fit your particular requirements, several commonly used types are described below.
Multiple Diameter Feeder This is a combination feeder and conveyor and the physical dimensions are variable on each. The small diameter feed end will operate at a full cross sectional load. When the material reaches the larger section, the cross sectional load will be at a controlled safe maximum.
Short Pitch Feeder This is also a combination feeder and conveyor. The short pitch end will handle full cross sectional loads. The material is then discharged into the standard section where the cross sectional load is reduced to the required maximum by the increase in screw pitch.
Variable Pitch Twin-Tapered Feeder This feeder is popularly used to unload bins or hoppers at a controlled rate. The feed opening under the bin is designed large enough to prevent material bridging and accepts materials uniformly across the length and width of the opening. This eliminates dead areas in the bin and reduces the chance of material bridging or spoiling.
Live Bottom Feeder Designed for use on straight sided bins, this feeder is composed of a number of feeder screws in tandem which serve as the bottom of the bin. Material is, therefore, drawn out equally from the full width. The Live Bottom Feeder is used to its best advantage on materials which tend to pack or bridge easily.
Screw Feeder Capacity The capacity table on pg. 20 shows Screw Feeder Capacities in cubic feet per hour per RPM. This table relates to full pitch or standard conveyors only. Shorter pitch flighting will convey a capacity in direct ratio to the capacity of the full pitch. For instance, a 9" conveyor with standard pitch (9") flighting on a 2-1/2" standard pipe, will convey 16.8 cu. ft./hr./RPM. The same conveyor, but with 3" pitch, will convey 1/3 this amount, or 5.6 cu. ft./hr./RPM. The capacity figure is theoretical. Actual capacity will often vary due to variation in head of material in the bin and variation in material characteristics.
Screw Feeder Speed The speed of the feeder screw can be determined by dividing the desired capacity in cu. ft. / hr. by the figure found in Capacity Table. For maximum efficiency, feeder screw speeds should be slower than standard screw conveyor speeds and allowances must be made for slippage of the material in the screw. Factors Affecting the Design Of A Screw Feeder 1. The material class The material physical characteristics 3. The capacity required 4. Material factor "F" Weight of material resting on the Feeder Screw 6. The dimensions of the feeder opening. 2.
5.
In designing a Screw Feeder, virtually every situation is unique in one respect or another. For this reason, we recommend that you consult KWS Engineering Department for proper recommendations concerning your particular needs.
Inclined Screw Conveyors Screw Conveyors can be operated with the flow of material inclined upward. When space allows, this is a very economical method of elevating and conveying. It is important to understand, however, that as the angle of inclination increases, the allowable capacity of a given unit rapidly decreases. A standard Screw Conveyor inclined 15° upward will carry 75% of its rated horizontal capacity. At an inclination of 25°, a standard conveyor may only handle 50% of its horizontal capacity. These are estimated figures and will vary with the characteristic of the material being handled. Inclined Screw Conveyor capacities can be increased over short distances, if no intermediate hangers are required. Other aids in conveying on an incline are the use of shorter than standard pitch and/or tubular housings or shrouded conveyor trough covers. Very often it becomes necessary to use high speed to overcome the tendency of material to fall back. The above aids are resorted to in order to overcome the tendency of a screw conveyor to become less efficient as the angle of incline increases. Vertical conveying by Screw Conveyor, on the other hand, is quite successful and it remains that a 45° incline or angles approaching this figure, are the most difficult on which to achieve successful conveying. Additional power is needed to convey on an incline. This added power is a function of the power required to lift the material. Judgment and experience in the art of conveying are required.
Tubular Trough and Half Pitch Screw Conveyor
CONVEYOR CAPACITY & SPEED CALCULATION: Visit the online engineering guide for assistance with using this calculator. DESIGN CONDITIONS
SPCL. FLIGHT TYPE - Place an "x" in the appropriate b
1. Flowrate(m) :
20000
2. Density :
45
3. Loading (K) :
30
lb/hr lb/Ft^3
%
7.
Cut Flight :
8.
Cut & Folded Flt.:
9.
Ribbon Type:
10. Paddles per Pitch:
FLIGHT PARAMETERS 4. Flight OD(D) :
in.
12
5. Flight ID(d) :
3.5
in.
6. Pitch(P) :
12
in.
CALCULATIONS : Shaftless Mass Flow Factor, C1
=
1.00
Spcl. Flight Type Factor, CF2
=
1.00
0
Flowrate (v)
=
m
=
444.4
Ft^3/Hr
=
12.93
Ft^3/Hr/rpm
Density
Capacity per revolution
=
0.7854 ( D² - d² ) P K 60 I C1 1728 CF2
Required Speed, N =
444.4
Ft^3/Hr
12.9
Ft^3/Hr/rpm
=
34.4
rpm
