Dear Shri Rekhawar,

While designing the conveyor, one has to fulfill the following conditions :

1) The belt tension should be adequate for transmission of torque at drive pulley/s.

2) The belt tension should be adequate so that sag value does not exceed the chosen design value.

The above conditions are to be satisfied during steady running, starting and stoppage (braking if applicable). One has to find out the belt tension at the location of gravity take-up in all the above cases. Then choose the value, which is the largest, and decide the take-up force. One can add some contingency as per his design practice. Such contingency could be 5 to 10% and need not be 50% or so.

The deciding of gravity take-up force for reversible conveyor or multiple drives would require some more application of mind.

Above is the general rule. If you are comparing the counterweight force with respect to only steady state condition, you may find it large. In case of relatively small length conveyor of low installed power, the tension requirement to limit the sag will be also dominant. In that case also one may feel that the counterweight is excessive with respect to steady state etc.

Regards,

Ishwar G Mulani.

Author of Book : Engineering Science and Application Design for Belt Conveyors.

Author of Book : Belt Feeder Design and Hopper Bin Silo

Advisor / Consultant for Bulk Material Handling System & Issues.

Pune, India.

Tel.: 0091 (0)20 25871916 ■

A. BELT SAG - Power & Deflection

Some design to 2% sag, as per some standards. Good design practice should hold the number to less than 1%. This is more critical for steel cord belt. The length of low tension sag zone must be considered. If long, then 1-2% sag will be a power hog. Such a condition is when belt has slight downhill slope.

B. INSTALLED POWER vs DEMAND - Design Criteria

Often designers wish to allow the operator to run the conveyor with installed power, not demand power. Since, operators will try to run with the maximum allowable tonnage regardless of design criteria. So, all structures, pulleys, take-up rating for running and acceleration/braking condition to prevent drive slip, all dynamics including control of sag (power, deflection, take-up motion,...) and belt vibration are also evaluated for installed power.

C. DRIVE TRACTION & TUP HYSTERESIS

In addition, the level of drive traction should be considered for nominal and likelyhood events. Thus, drive friction may have multiple values depending on weather et al. Part of this equation may include hysteresis loss from routing of TUP wire rope and sheaves which may need compensation to meet design criteria.

D. DYNAMIC ANALYSIS & BELT ACCUMULATION @ INFLECTION PT.

If the conveyor has a relatively long horizontal run and then rises, belt will gather at the inflection point, pulling belt from the TUP. This occurs when the horizontal run coasting time will be greater than the incline stopping time. Often, the designers software gives a hint of this condition. However, the software should also have belt gathering analysis between idlers (belt sag) that shows the sensitivity during a coast-to-stop.

E. GOOD DESIGN PRACTICE

So, what is required for good design. This may differ according to designers experience with the product, local conditions, .... Taking information from a catalog or standard is not best practice. ■

Good morning Rekhawar..

The answer to your question could be that one should calculate the value of the slack side tension based on mu of 0.35 as applied to Te start.

If you apply 0.35 to Te, as David says, you can get slip during starting.

Since Te start = say Te x 1.4, then there is the 40% you are looking for.

Cheers

LSL Tekpro ■

Indeed, it does give a reasonable result David.

The beauty of the Te x start factor approach is that one can select the correct start factor for each application.

For example I use 1.2, 1.4, and 2 for VSD, delayed fluid coupling, and direct coupled respectively.

Cheers

LSL Tekpro ■

All good interesting stuff Dave..

What do we learn from it all?... well..

- Design T2 your way and there will be no drive slip.

- Design T2 my way, and there will also be no drive slip.

(I base this on the fact that I cannot ever recall any of my conveyors ever having drive slip over the last 30 years.)

- Design T2 only on running Te and mu of.35 and you are at risk.

I rest my case your honour..

Cheers Dave

LSL Tekpro ■

Dave,

I take issue with your out-of-hand claim that a belt-pulley contact friction must be < f=0.35 in order to work in all condition otherwise you will experience slip.

First, the problem is more complex than you or Graham claim. The factors used are old "cover your rump" numbers. Both of you are better engineers and have far greater experience than these simple assumptions demonstrate.

Please note, many conveyors have done well with the published numbers, considered by many to be standard and common practice. There is no need to cause the removal of the standard values as they apply to the many common systems simple systems for most products and light duty environments, dry and wet .

Second, there are specific conditions which do require special attention and, which in some cases, you have noted and some you have not. A little traveling music of some of the many contributions are:

1. Rubber glazing and ageing of bottom belt cover and lagging cover

2. Loss of water grooves to aid removal of moisture in contact zone

3. Slime contamination from iron, bauxite, nickel and other fine material than has little shear strength under the contact pressure

4. Reduced arc of contact due to stiff belts and shortened idler trough transition zones - Germans used to apply a 5 degree contact arc penalty when designing higher strength systems

5. Load sharing inaccuracies between adjacent drives

6. Too close proximity of non-driven pulleys that cause contact irregularities.

7. Heavy rains and belt speed which require faster than normal removal of moisture at the point of contact

8. Downhill conveyors where moisture enters at T2 which has less squeegy action

9. Ice, snow and frost conditions that obtain access to drive-pulley interface

10. Apply the work of Zedde's Doctor of Engineering Thesis on drive contact friction.

11. Geometry of the lagging, lagging grooves, bumpy surfaces, other surfaces, or not , for rubbers, ceramics; others

12. Material of lagging - some rubbers have greater grip, some ceramics have less than expected, some ceramics use different principles to create friciton

The point to this pedantic ranting is: when designing conveyors of consequence, some level of experience and know-how should be included in the procedure to “fit-for-purpose” engineering. The artificial number "coefficient of friction" is not single valued. It should be viewed as part of the cost-benefit or risk-reward HAZOP conditional analysis, with many factors, and should be associated with knowledge of the complexities.

These comments are not.comprehensive, just a word of caution. ■

Hmmmmm,

I would have to say - design with caution and use expertise advise

We have a yard conveyor - 500m long..2000tph with travelling stacker, chevron stacking for 2 piles

GTU at tail pulley, cermaric drive pulley with close located snubber pulley.

Belt is fabric 5ply.

When running, there is some edge buckling at the bend point up the tripper car - no spillage.

We have had a lot of wet weather in CQ of late.

We actually broke / snapped the belt in half - straight across break.

The drive is 132kW with delayed fill turbo coupling.

Loking at the belt - we could see "slip" marks on the cermamic nibs in the return side and then the normal "indent" marks. These aligned up with the linear dims back to the drive pulley and the break. The break occured on a loaded re start.

I am assuming that;

The GTU system did not keep T2 high enough for the start - the GTU is a horizontal travelling trolley - can be issue with buildup around the wheels

The loss of T2 led to drive spillage

The restoration of T2 and stoppage of slip led to a very high T1 jerk.

Belt broke (belt is unbranded - must be OS supply from Oz).

I would not think that I would see the day a belt break straight across.

Yes - Ithink that T2 and GTU selection is important and simply use of mu can be dangerous.

Cheers

James ■

Hi David

Your comment wrt FC delay start time is valid....I have not checked this.

I share your views of FC use and appoications - 140% max FLT is based for one set of conditions...we have a travelling stacker..so actual load cases will vary with stacker position and direction of travel.

The tail end GTU unit is prob. the best wrt geometry for the yard conveyor but due checks have to be made to make sure it works.

I have never seen a belt break before like this - was not even at the splice.

Message is that this can happen if the system is not well designed considering all the possible scenarios.

Cheers

James ■

Interesting exchange of views and information. However, I am a little bit puzzled by statement by Mr. David Beckley regarding DIN standard. As far as I can see DIN standard is very specific in the need to analyse both stedy state and non - steady cases of operation and accordingly requires calculation of minimum traction forces. It is, however, true that frequently ( at least in South Africa) designs tend to ignore start up and stopping in sizing of take up parameters.

Going back in years I am aware of various design recommendations which aimed at compensating for the issues described by Mr. Nordell but "unaccounted" in the available standards. This is an interesting aspect in itself where a gap exists between knowledge gained from research and common, recommended (?) practice.

Regards

Marian Otrebski ■

Other errors in defining the true take-up (TUP) force also lead to errors in defining the true slack side tension T2.

Non-vertical TUP systems often have trolleys on rails and wire rope and sheaves of varying complexity. These contribute drag against trolley TUP motion, which can exacerbate the degree of TUP force. Sheaves often use bushings which further add to the loses. Bending of wire rope around sheaves adds to the loses. ■

Larry:

"Bending of wire rope around sheaves adds to the losses"

You know, once when I extended the Namakwa Sands Dual Carry Conveyor, I needed to adjust the take-up mass a little.

It was supposed to be 11 tons. The crane scale I used to weigh it indicated 14 tons.

(I can assure you, I nearly soiled my silk underwear).. So...

I removed the counterweight, I weighed a truck, I put the counterweight on the truck and weighed it again.

Actual weight 11.2 tons.

I wonder how many people guestimate take-up masses by working out the volume and densities, and confirm with crane scales, which are about as accurate as Robert Mugabe at trying to tell the truth.

The effect of bending of wire rope around sheaves adding to the losses, is as negligible as adding the last of George Bushe's brain cells to the take-up mass.

Cheers

LSL Tekpro ■

Graham,

TUP hysteresis losses in sheaves and ropes are dependent on rope construction, tension, and size of sheave to rope diameters, so says the US and German governments.

Aside from the many projects we have worked on, it became a big concern in Hanford, Washington nuclear power plant crane lift losses. A USA government study was conducted over a number of years and at great expense, which developed the analytic loss factors. Hundreds (>200) of tests were conducted that quantified the mechanisms at work in the losses. Their main interest was in the potential rope failure in fatigue and the contributors.

I have that report and have used the Hanford laboratory equipment to measure loss factors in ropes with respect to belt splices. Belt splices are also influenced by their cord diameter to pulley diameter. Why are pulleys sized to the steel cord diameter? Bridgestone also made a fatigue study on their special rope designs.

Germany also conducted research into rope losses over sheaves. Their work was also published together with the theory behind the losses. Their work was made a part of the Hanford study. A company in Los Angeles, Tension Member Technology, purchased the equipment from Hanford. They do a successful business into ropes and sheaves failure mechanisms in steel and non-steel ropes. The owner was the principle party at Hanford.

There is a mine in Chalice Idaho where we were called in to assess the TUP pulley shaft failure. If part, the TUP pulley loads where larger than the designer had predicted which contributed to the shaft fatigue break. There were two gravity TUP towers, on each side to split load. Special load cells were placed at the counterweight and at the trolley. We measured 44% losses in the sheave and ropes Did the same thing on a conveyor at the Syncrude Oil Sands. The same method was applied to 9km downhill at El Abra in Chile and so on.

Disbelieve at your own expense. I will kndly decline to visit your many successful installation for fear of catching a cold. ■

In the Chalice sudy, the losses where measured with conveyor in operation to determine the hysteresis losses during variations in the TUP movement and TUP tension function.

We determined the influence on drive traction (T2) and on all pulley forces with their structural loads that where impacted by the large hysteresis losses.

TNT have better than a dozen different machines to measure rope and sheave fatigue interaction up to 200 mm rope diameter.

There is a lab in Scotland that also performs tests on larger ropes and their hysteresis and fatigue properties in the dynamic state.

Read any major rope house design catalog, such as Haggie Rand in RSA, which apply these rope/sheave losses and fatigue to the deep mine hoists.

Haggie Rand Ltd: P.O. Box 52: Germinston, South Africa: 1400. Phone: 27-11-825 1076: FAX: 27-11-873 9389. ■

Hi Larry..

"We measured 44% losses in the sheave and ropes"

One of the pre-requesites for the take-up rope system is that the sheaves must actually rotate, and be blessed with a suitable dose of Shell Alvania 3 all purpose grease for good measure.

(Failing that, next time you are in South Africa, I could help you get your sheave loss meter fixed.)

It is hot summer here at present, I'm sure you wouldn't catch cold should you visit, and I would be delighted to show you some of our installations.

I'm sure you would be very impressed.

Cheers

LSL Tekpro ■

Wishing one and all well.

Rather than carry the torch any further I suggest a little bed time storybook from your favorite wire rope supplier. You guys need to become informed of what the rope suppliers publish on rope bending on sheaves drag losses separate from sheave drag. From memory its about 4% per sheave, not including the bearing drag. This assumes a rope to sheave ration acceptable to the rope manufacturer.

I like the idea of visiting the esteemed systems of South Africa that have eluded my radar. Since I see Graham wandering Australia in search of an idea or two, turnabout is fair play. ■