Koepe/Friction Hoists

1. Introduction

The friction (or Koepe) hoist is a machine where one or more ropes pass over the drum from one conveyance to another, or from a conveyance to a counterweight. In either case, separate tail ropes are looped in the shaft and connected to the bottom of each conveyance or counterweight. The use of tail ropes lessens the out-of-balance load and hence the peak horsepower required of the hoist drive. When compared with a drum hoist for the same service, the tail ropes reduce the required motor HP rating by about 30%, but the power consumption remains virtually the same. Tail ropes have been used for a few double-drum hoist installations to the same effect, but this practice has not gained acceptance by the mining industry.

Because they normally use several hoisting ropes, the largest friction hoists can handle heavier payloads than the largest drum hoists. The drum hoists are normally limited to the capacity of a single rope. Friction hoists require a higher safety factor (SF) on the hoist (head) ropes and are not considered practical for really deep shafts employing high rope speeds.

For mine applications, Koepe hoists compete with drum hoists and the decision concerning which one is best suited for a particular application is considered in the example presented as a side study in Chapter 6 – Feasibility Studies. Drum hoists are discussed separately in Chapter 13. The hoist cycle times developed for drum hoists in that chapter have equal application to friction hoists.

For historical reasons, friction hoists (unlike drum hoists) are usually thought of in terms of metric rather than imperial (British) units. To describe the size of a friction hoist people will say “a 3m wheel diameter” rather than “a 10-foot hoist.” For this reason, the explanations and design calculations that follow are mainly performed in metric units of measure.

2. Rules of Thumb

Hoisting Distance

• A friction hoist with two skips in balance is normally suitable for hoisting from only one loading pocket horizon and for a hoisting distance exceeding 600m (2,000 feet). Otherwise, a counter-balanced friction hoist (conveyance and counterweight) is usually employed (for multilevel, shallow lifts, or cage hoisting). Source: Ingersoll-Rand

• The practical operating depth limit for a friction hoist is 1,700m (5,600 feet) for balanced hoisting and 2,000m (6,600 feet) for counterweight hoisting. Beyond these depths, rope life may be an expensive problem. Source: Jack de la Vergne

• The hoisting ropes (head ropes) for a friction hoist are not required to be non-rotating for depths of hoisting less than 800m (2,600 feet) provided right hand and left hand lays are employed to cancel rope torque effect. Tail ropes must always be non-rotating construction and connected with swivels at each end. Various sources 

Static Tension Ratio

• For a tower mounted skip hoist, the calculated static tension ratio (T1/T2) should not exceed 1:1.42, but 1:1.40 is preferable. 

For a ground mounted skip hoist, the calculated static tension ratio should not exceed 1:1.44 but 1:1.42 is preferable. For a cage hoist installation, these values may be exceeded for occasional heavy payloads of material or equipment transported at reduced speed. Various Sources

Tread Pressure

• Tread pressure should not exceed 17.5 kg/cm2 (250 psi) for stranded ropes and 28 kg/cm2 (400 psi) for locked coil ropes. Source: A.G. Gent

• For lock coil hoist ropes, the tread pressure calculated for skip hoists should not exceed 2,400 kPa (350 psi), or 2,750 kPa (400 psi) for a cage hoist when considering occasional heavy payloads of material or equipment. Source: Jack de la Vergne

• For stranded hoist ropes, the tread pressure calculated for skip hoists should not exceed 1,700 kPa (250 psi) or 2,000kPa (275 psi) for a cage hoist when considering occasional heavy payloads of material or equipment. Source: Largo Albert

Tail Ropes

• The natural loop diameter of the tail ropes should be equal to or slightly smaller than the compartment centres. Source: George Delorme

Hoist Wheel Rotation

• The total number of friction hoist wheel revolutions for one trip should be less than 100 for skip hoists, but may be as high as 140 for cage hoists. Source: Wire Rope Industries and others

• The hoist wheel rotation at full speed should not exceed 75 RPM for a geared drive, or 100-RPM for a direct drive. Source: Ingersoll-Rand

Position

• The distance between the hoist wheel and the highest position of the conveyance in the headframe should not be less than 1.5% of the distance from the hoist wheel to the conveyance at the lowest point of travel. Source: Largo Albert

• At full speed, a time increment of at least ½ a second should exist as any one section of rope leaves the hoist wheel before experiencing the reverse bend at the deflector sheave. Source: George Delorme

• The clearance between the bottom of the conveyance, at the lowest normal stopping destination in the shaft, and the top of the shaft bottom arrester (first obstruction) is usually 5 feet. This arrangement ensures that the weight of the descending conveyance is removed from the hoist ropes. Source: Largo Albert

• The tail rope loop dividers are generally placed below the arrester. The bottoms of the tail rope loops are then positioned 10 to 15 feet below the dividers. Beneath this, a clearance of about 10 feet will allow for rope stretch, etc. Source: Largo Albert

Hoist Speed

• Where the hoist line speed exceeds 15m/s (3,000 fpm), the static load range of the head ropes should not be more than 11.5% of their combined rope breaking strength. The (ratio of) hoist wheel diameter to rope (stranded or lock coil) diameter should not be less than 100:1, and the deflection sheave diameter to rope diameter should not be less than 120:1. Source: E J Wainright

• The maximum desirable speed for a friction hoist is 18m/s (3,600 fpm). Source: Jack Morris

• The maximum attainable speed for a friction hoist that can be safely obtained with today’s (1999) technology is 19m/s (3,800 fpm). Source: Gus Suchard

• In North America, the desirable speed for cage service is approximately 2/3 of the optimum speed calculated for a skip hoist for the same hoisting distance. Source: Jack de la Vergne

Hoist Wheel Specifications

• The hoist wheel diameter to rope (lock coil) diameter should not be less than 100:1 for ropes up to 1-inch diameter, 110:1 for ropes to 1½ inches diameter, and 120:1 for ropes to 2 inches diameter. Source: Glen McGregor

• A ratio of 100:1 (wheel diameter to lock coil rope diameter) is adequate for ropes of 25-35 mm diameter. This should increase to 125:1 for ropes of 50-60 mm diameter. Source: Jack Morris

• Rope tread liners on the hoist wheel should be grooved to a depth equal to one-third (1/3) of the rope diameter when originally installed or replaced. The replacement (discard) criterion is wear to the point that there is only 10 mm (3/8 inch) of tread material remaining, measured at the root of the rope groove. Source: ASEA (now ABB)

• On most fiction hoist installations, the maximum tolerable groove discrepancy is 0.004 inches, as measured from collar to collar. Source: Largo Albert

Production Availability

• A friction hoist is available for production for 108 hours per week. This assumes the hoist is manned 24 hours per day, seven days per week, and that muck is available for hoisting. Source: Jack Morris

• With proper maintenance planning, a friction hoist should be available 126 hours per week (18 hours per day). Source: Largo Albert

Spacing

• The minimum distance (design clearance) between a rope and bunton or divider is 5 to 6 inches. This is mainly because the hoist rope vibration is normally 2 to 3 inches off centre; 4 inches is considered excessive. Source: Humphrey Dean

• The spacing between head ropes should be 1 inch for each foot diameter of the hoist wheel to get an adequate boss for the deflection sheave. Source: Gerald Tiley

3. Tricks of the Trade

• The easy way to design a friction hoist is to first determine the required hoisting speed and payload then determine the ropes that are needed to meet the required SF. The hoist parameters can then all be determined only considering the hoist ropes and line speed. Source: Tom Harvey

• The distinguishing feature that should be recalled when designing or operating a friction hoist is that “weight is your friend.” In other words, heavier ropes and suspended loads mean higher force of friction and greater facility for braking, etc. Source: Richard McIvor

• The rule of thumb (attributed to Wainright) that indicates a minimum SF of 7 for friction hoist head ropes is not correct. There are a very large number of hoist installations worldwide that have operated satisfactorily for many years at smaller SFs. In this respect, the regulations stipulated for the Province of Ontario in Canada are a good guideline, anywhere. Source: Largo Albert

• To avoid stress concentrations, it is desirable to manufacture a friction hoist wheel in one piece. Wheels up to about 3m (10 feet) in diameter can be shipped complete with shaft to most locations. Source: Gerald Tiley

• When designing a tower-mounted friction hoist, consideration should be given to the possible avoidance of deflection sheaves, as they represent a maintenance headache. Source: Richard McIvor

• At full speed, a time increment of 0.6 second should exist as any one section of rope leaves the hoist wheel before experiencing the reverse bend at the deflector sheave. This adds to the headframe height, but the added clearance is desirable for maintenance and change-out of the sheave wheels. Source: Largo Albert

• While it is better to have the rope spacing the same at the hoist wheel and the head sheaves for a ground mounted Koepe hoist, this is not necessary provided that the fleet angle of the outside ropes is 10 or less. This is known because there are single rope friction hoists in Europe with both head sheaves on the same headframe deck that operate satisfactorily, provided the fleet angle is maintained at 60 minutes (10) or less. Source: Tréfileurope

• For a single rope ground mounted Koepe hoist, it is better to have the head sheaves in the same plane as the hoist wheel. However, the head sheaves may be mounted on the same deck of the headframe tower, provided the fleet angle of the outside ropes is not more than 1½ to 2 degrees. Source: Henry Broughton

• While it is better to have the rope spacing the same at the hoist wheel and the skip attachment, this is not necessary provided the fleet angle of the outside ropes is 10 or less when the conveyance is at its upper end of travel. Source: Borje Fredricksson

• The arresters (“last resort”) at the shaft bottom are designed to stop a full-speed conveyance at 2g, while an ascending conveyance must be stopped at less than 1g (i.e. 0.9g), although not necessarily from full speed if it exceeds 15m/s (3,000 fpm). Various sources

• The tail ropes should be oriented to overcome the Coriolis effect. If placed in the East-West direction, the tail ropes will freely open and close. If the compartments are North and South, the ropes will foul the separating spacers (loop dividers) if not widely spaced. Source: Gerald Tiley

• The Coriolis effect can be neglected, as it is much smaller than the movement at acceleration/deceleration and due to rope torque of the tail ropes. Source: Borje Fredricksson

• High-speed friction hoists [over 12m/s (2,400 fpm)] are oriented with the wheel diameter East-West to minimize the effect of Coriolis acceleration on the tail ropes. Source: Jack Morris

• The effect of Coriolis acceleration on the tail ropes is diminished when a fixed guide system is employed, as opposed to using rope guides. Source: Jack de la Vergne

• The tail rope weight is normally designed equal to the head ropes; however, tail ropes slightly heavier than the head ropes will assist acceleration from the loading pocket. Slightly lighter tail ropes will provide a greater SF for the head rope section above the conveyance as it approaches the highest point of travel (the point at which uneven rope tension is most severe). Source: Gerald Tiley

• The distance between head ropes (spacing) varies between 8 inches and 12 inches. At 8 inches, some installations experience rope slap but this is not considered a serious problem, since the ropes are running at the same speed. (Author note: regular slapping is said by others to lead to martensitic alteration, resulting in broken wires.) Narrow rope spacing may require that the rope attachments at the conveyance be staggered. This can be accomplished by including a link at every other attachment. Source: Humphrey Dean

• The guideline for rope spacing is 8 inches up to 11/8-inch rope diameter, 10 inches to 1¼ inches, 12 inches to 1½, 14 inches to 15/8, and 16 inches to 1¾. Drawhead connections can be staggered but this is costly and complicates rope adjustment and maintenance. Source: Largo Albert

4. Friction Hoist Design

Listed below are the steps in designing and selecting a friction hoist.

1. Determine the SF required for the given hoist distance

2. Determine hoisting speed, V

3. Calculate the hoist cycle

4. Define (cage) or calculate (skip) the payload

5. Determine the weights of the conveyances required

6. Select hoist (head) ropes

7. Determine the wheel diameter of the hoist

8. Select balance (tail) ropes

9. Calculate the RMS power requirement

Example

Design and select two friction hoists at the same time. One is required for production hoisting and the other for cage service.

Facts: 

1. Both hoists will be tower mounted in the same headframe

2. The skip hoist requires a capacity of 500 tonnes/hour

3. The cage hoist requires a payload of 26 tonnes

4. Each has a hoisting distance, H of 1,000m

5. The statutory SFs of Ontario, Canada are to apply

Solution:

Step 1: Determine the SF required for the given hoist distance.

Following is the SF required by statute for the hoist ropes.

• SF = 8 -.00164D in which D = length of suspended rope, hence D= approximately H + 50m to account for rope suspended above the dump and beneath the loading pocket (in the case of a skip hoist) and similar extra rope length in the case of a cage hoist.

• SF = 8 - (.00164 x 1,050) = 6.3

Step 2: Determine hoisting speed, V.

• The optimum skip hoisting speed, V = 0.44 H ½ = 14m/s (rounded)

• A suitable cage hoisting speed will be about 2/3 V =10m/s (rounded up)

Step 3: Calculate the hoist cycle.

Calculate the skip hoist cycle time, T. (Since the hoisting distance exceeds 600m, balanced hoisting with two skips is determined.)

• T = H/V + 1.3 V + 25 = 115 seconds

• Trips per hour = 3,600/115 = 31.3

Calculate the cage hoist cycle time (a cage and counterweight is assumed).

• T = 2H/Vc + 2.6 Vc + 70 = 296 seconds

• Trips per hour = 3,600/296 = 12.1

Step 4: Define (cage) and calculate (skip) payload.

The cage payload is given at 26 tonnes and the skip payload is calculated by dividing the capacity per hour by the trips per hour.

• Skip payload, P = 500/31.3 = 16 tonnes (rounded up)

Step 5: Determine the weights of the conveyances required.

The weight of the cage for 26-tonne capacity will be approximately 20 tonnes if it is steel (steel is typical for friction hoists). The weight of the counterweight will be made equal to the empty cage weight plus half the payload = 33 tonnes. (In this case, the counterweight will be designed to readily remove a portion of its weight for regular cage service with lighter payloads.) The weight of the skip, S will be approximately 13 tonnes (refer to Table 15.5a, Chapter 15 of this handbook) for a steel bottom dump skip that would normally be used for this application. However, this weight might not be enough to maintain the required tension ratio (in this case, the skip would be “ballasted” with extra weight).

The empty weight of skip required to maintain a tension ratio of 1.40:1 follows.

• St = P{2.5 - (H x SF/4,500)} = 16 (2.5 - 1.4) = 17.6 tonnes

• The skips will be ballasted to weigh 17.6 tonnes

Step 6: Select hoist (head) ropes.

Cage Hoist

Try 6 lock coil ropes of 32 mm diameter weighing 5.58 kg/m and having a breaking strength (BS) of 890 kN.

• SF obtained = Number ropes x BS/maximum suspended load

= 6 x BS/weight of ropes, payload and cage

= 6 x 890/g (35.5 + 26 + 20) = 6.7 (6.3 required)

• T1/T2 obtained = (35.5 + 26 + 20)/(35.5 + 33) = 1.19

• T2/T3 obtained = (35.5 +33)/(35.5 + 20) = 1.23

Maximum total suspended load = (35.5 + 26 + 20) + (35.5 + 33) = 150 tonnes

Skip Hoist

Try four of the same lock coil ropes – 32 mm diameter, 5.58 kg/m, and BS 863 kN.

• T1/T2 obtained = (23.7 + 16 + 17.6)/(23.7 + 17.6) = 1.39 (1.40, or less, desired) Maximum total suspended load = (23.7 + 16 + 17.6) + (23.7 + 17.6) = 98.6 tonnes

• SF obtained = Number ropes x BS/maximum suspended load

= 4 x BS/weight of ropes, payload and skip

= (4 x 890/g(23.6 + 16 + 17.6) = 6.35 (6.3 required)

Step 7: Determine the wheel diameter of the hoist, D.

The statutory requirement for lock coil ropes is 100 times the diameter of the hoist rope at this location.

Cage and Skip Hoist

(Statutory) D = 100d = 100 x 32 = 3,200 mm = 3.2m

The statutory diameter may not be sufficient. For example, the diameter should be increased if the permitted tread pressure is exceeded, the number of wheel revolutions per trip is too high, or the ropes are greater than 35-mm diameter.

The tread pressure is calculated by dividing the total suspended load by the projected contact area of the ropes on the hoist wheel. Tread pressure should not exceed 2,400 Kpa for a skip hoist or 2,750 kPa for a cage hoist with maximum payload.

• Cage hoist tread pressure = 150g x1,000/ (6 x 3.2 x 32) = 2394 kPa (V)

• Cage hoist revolutions = H/πD = 1,000/3.2π = 99.5 (V)

• Skip hoist tread pressure = 98.6g x1,000/ (4 x 3.2 x 32) = 2360 kPa (V)

• Skip hoist revolutions = H/πD = 1,000/3.2π = 99.5 (V)

A wheel diameter of 3.2m should be satisfactory for both hoists. (On detailed investigation, it may be increased to 3.5m to increase rope life).

Step 8: Select balance (tail) ropes.

Select non-rotating balance (tail) ropes matching the head rope weight and with a natural loop diameter equal to the compartment spacing.

Note

Tail ropes can be custom manufactured to meet precise weight requirements (i.e. kg/m). For the cage hoist (assuming no deflection sheave), select three non-rotating ropes weighing twice the head rope weight. The head ropes weigh 5.58 kg/m; therefore, the tail ropes will weigh 11.16 kg/m with a 53 mm diameter. If the ropes are 34 by 7, the natural loop diameter will be 46 x 53 = 2,438 mm (unsatisfactory). If the ropes are 18 by 7, the natural loop diameter will be 60 x 53 = 3,180 mm (satisfactory).

For the skip hoist (assuming a deflection sheave is required to bring the conveyances closer together in the shaft, say 2m between compartment centres), select three nonrotating ropes of weight = 4/3 x 5.58 = 7.44 kg/m with a 43 mm diameter. If they are 34 by 7, the natural loop diameter will be 46 x 43 = 1,978 mm (satisfactory).

Step 9: Calculate the RMS power requirement.

Assume there is a force-ventilated DC or cyclo-converter drive. 

The skip hoist RMS power = a constant(k) x unit weight of the ropes x (speed)5/4

SF obtained

• k=24 for a standard DC (FV) or cyclo-converter drive (FV)

• The skip hoist RMS power = (24/6.3) x 4 x 5.58 x 141.25 = 2,300 kW (3,100HP)

• The cage hoist RMS power = skip hoist factored for speed and out-of-balance loads =

2,300 x 10/16 x (10/14) 1.25 = 944 kW (1250) HP)

5. Production Availability

Confusion and controversy exists in the mining industry as to the meaning of the word “availability” when applied to mine hoists. For hoist maintenance personnel, it may mean the percent of the time the piece of equipment is available to work compared with the total time available. On the other hand, those engaged in selecting and evaluating hoists for mine service must consider the availability of the total hoist system, taking not only maintenance downtime into account, but also downtime due to shaft repairs, power outages, rope dressing, skip change-out, etc. This chapter is concerned with the availability of the total system, and for this purpose, it is described as “production availability.”

To determine the production availability of a friction hoist for the purpose of estimating hoisting capacity per day, a detailed calculation should be made for each case, taking into account the total hoisting system. This will include allowances for empty loading pocket, full bin, hoisting spill, etc. Availability will usually be slightly less than a drum hoist because the friction hoist demands a more sophisticated maintenance routine. It is higher for a five or six days per week operation (because some maintenance work can be performed on the weekend) than a seven days per week operating scenario.

Example

Facts: 

1. Estimate is based on a 7-day workweek

2. Automatic hoisting is assumed

3. Two skips are hoisted in balance

4. No cage service is required

5. 12-day annual shutdown is assumed

Solution:

The hoist plant availability is shown in Table 14-1.

Table 14-1 Hoist Plant Availability

(Friction (Koepe) Hoist – Seven Day per Week Operation)

6. Comparisons

Following is a comparison of ground versus tower mount friction hoists.

Ground Mount Friction Hoist

Listed below are ground mount friction hoist advantages.

• Shorter headframe.

• Steel headframe (concrete is preferred in tower mounts for rigidity – reinforced concrete is not subject to residual stresses).

• An elevator is not required in the headframe.

• An overhead bridge crane may not be required.

• Easier access for maintenance.

• A heated headframe is not required.

• A water supply to the top of the headframe is not required.

• Shorter runs of power cables.

• Less susceptible to damage from overwinds, mine explosions, lightning, and earthquakes.

• The longer rope between the hoist and the highest point of conveyance travel makes rope surge and possible subsequent structural upset less likely.

• Most efficient use of available space in the shaft for conveyances.

• Ability to operate without problems at a higher tension ratio (T1/T2). This is likely due to the dampening effect obtained from the wide-angle wrap of hoist rope around the head sheaves and the greater distance between the high point of travel for the conveyance and the hoist wheel.

Tower Mount Friction Hoist

Listed below are tower mount friction hoist advantages.

• Zero or one deflection sheave is required. Two are required for a ground mount – one is subject to reverse bending of the hoist ropes.

• Installing and changing head ropes is less complicated.

• Less real estate is occupied.

• The hoist ropes are not subject to the elements – icing is less of a concern.

• Rope vibration (whip) is less of a concern.

• The headframe tower may be more aesthetically pleasing.

• The headframe shell can be used for shaft sinking simultaneous with Koepe hoist installation above the sinking sheave deck.

WAKTU PEMBENTUKAN MINYAK DAN GASBUMI

konsepsi stadium serpih. Waktu pembentukan minyak - gasbumi sangat erat hubungannya dengan mekanisme transformasi dan mekanisme migrasi. Juga hal ini sangat erat hubungannya dengan terjadinya suatu akumulasi minyak, ada tidaknya suatu perangkap pada wakttu minyak dikeluarkan.
Pada umumnya ada 2 tanggapan mengenai waktu pembentukan ini :
pembentukan segera ( early formation )
pembentukan lambat ( late formation ).
ANGGAPAN PEMBENTUKAN SEGERA
anggapan ini didasarkan pada banyak hal :
- terdapatnya hidrokarbon dalam sedimen resen. Smith, 1952 (di gulf of Mexico,1954 ; Santa Crus Basin), menunjukan bahwa minyak bumi dapat terbentuk tak lama setelah sedimentasi. Hal yg sama ditemukan oleh Kidwel dan Hunt (1958), malahan akumulasi daqat terjadi dalam waktu beberapa puluh ribu tahun saja, seperti di Pederneles, Venezuela.
- kenyataan bahwa makin tertimbun sedimen, lempung dan serpih makin padat sehingga makin sulit bagi cairan yg terbentuk didalamnya untuk bermigrasi. Kenyataan ini dikemukakan oleh Hedberg (1932) dengan fasa perkembngan yg didasarkan atas percobaan inti pemboran didaerah venezuela.

Drum Hoists

1. Introduction

Drum hoists are employed in mines on tuggers, slushers, stage winches, cranes, rope tensioners, and even for long plumb line winches. Chapter 13 is mainly devoted to drum hoists that serve as mine hoists (winders). These machines are the most significant hoists in a mine, used for hoisting the ore and waste rock as well as moving personnel, equipment, and materials into and out of the mine.

Single-drum mine hoists are satisfactory for limited application; however, most are manufactured double drum to facilitate balanced hoisting of two conveyances in the shaft. Balanced hoisting can be accomplished with a single-drum hoist for shallow applications that require a single layer of rope on the drum. In this case, the rope being wound is wrapped in the same grooves that are vacated by the other rope being unwound. Single-drum hoists have been built with a divider flange and even with the drums of different diameter on either side of the flange (“split-differential”) to accomplish balanced hoisting – these designs are no longer manufactured. This chapter is largely devoted to the double-drum mine hoist because it is by far the most common type employed.

All mine hoists manufactured today are driven electrically by motors that have an independent ventilation source. Having an independent source reduces the horsepower requirements by more efficient cooling of the windings especially during slow-speed operations and permits filtering of the air that reaches the motor.

Until recently, DC drives with solid-state converters (thyristors) were almost exclusively employed. Lately, larger mine hoists have been manufactured with AC drives that are frequency controlled (cyclo-converter). Typically, the larger double drum hoists are direct driven with overhung armatures, while double helical open gears drive those of medium size. Small or very slow mine hoists may employ a gearbox for speed reduction.

For mine applications, drum hoists compete with friction hoists (refer to Chapter 14). The decision concerning which one is best employed for a particular application is discussed as an example of a side study in Chapter 6 – Feasibility Studies. Some hoisting parameters explained in this chapter (e.g. hoist cycle time) have equal application to friction hoists.

For historical reasons, drum hoists (unlike friction hoists) are still thought of in terms of Imperial rather than metric units. To describe the size of a drum hoist, miners will say “a 10-foot hoist” rather than “a 3m hoist.” For this reason, the explanations and calculations that follow are mainly performed in Imperial units. The Blair multi-rope (BMR) hoist (a variation of the double-drum hoist) employed for extremely deep shafts is not discussed in this Chapter.

2. Rules of Thumb

Hoist Speed

• The maximum desirable speed for a double-drum hoist with fixed steel guides in the shaft is 18m/s (3,600 fpm). Source: Peter Collins

• The maximum desirable speed for a drum hoist with wood guides in the shaft is 12m/s (2,400 fpm). Source: Don Purdie

• An analysis of the theory developed by ASEA (now ABB) leads to the conclusion that the optimum speed is a direct function of the square root of the hoisting distance. Applying the guideline of 50% and assuming reasonable values for acceleration and retardation leads to the following rule of thumb equation for the optimum economic speed for drum hoists, in which H is the hoisting distance.

Optimum Speed (fpm) = 44H½ , where H is in feet

Or, Optimum Speed (m/s) = 0.405 H½ , where H is in metres

Source: Larry Cooper

• Assuming reasonable values for acceleration gives the following rule of thumb equations for the design speed of drum hoists, in which H is the hoisting distance (feet).

Design Speed (fpm) = 34 H½ , hoisting distance less than 1,500 feet

Design Speed (fpm) = 47 H½ , hoisting distance more than 1,500 feet

Source: Ingersoll-Rand

• The hoist wheel rotation at full speed should not exceed 75 revolutions per minute (RPM) for a geared drive, nor 100-RPM for a direct drive. Source: Ingersoll-Rand

Hoist Availability

• With proper maintenance planning, a drum hoist should be available 19 hours per day for a surface installation, 18 for an internal shaft (winze). Source: Alex Cameron

• A drum hoist is available for production for 120 hours per week. This assumes the hoist is manned 24 hours per day, 7 days per week, and that muck is available for hoisting. Source: Jack Morris

• The total operating time scheduled during planning stages should not exceed 70% of the total operating time available, that is 16.8 hours per day of twenty-four hours. Source: Tom Harvey

• In certain exceptionally well-organized shafts, utilization factors as high as 92% have been reported, but a more reasonable figure of 70% should be adopted. With multi-purpose (skipping and caging) hoists, the availability will be much lower. Source: Fred Edwards

Rope Pull

• The manufacturer’s certified rope pull rating for a drum hoist assumes the rope flight angle is 25 degrees or more from the horizontal. The rope pull rating should be reduced by 10% for an installation where the ropes run horizontally between the hoist and the head sheave. Source: Ingersoll-Rand

Hoist Drums

• The hoist drum should be designed to coil rope for the hoisting distance plus an allowance equal in length to 10 dead wraps on the drum. Source: John Stephenson

• The hoist drum should be designed to coil sufficient rope for the hoisting distance plus an allowance of 500 feet, for most applications. Very deep shafts may need 600 feet of allowance. Source: Jack de la Vergne

• The hoist drum should be designed to coil sufficient rope for the hoisting distance plus the statutory three dead wraps, the allowance for rope cuts and drum pull-ins for the life of the ropes plus at least 200 feet of spare rope. (At least 250 feet of spare rope is desirable for deep shafts.) Source: Largo Albert

• The pitch distance between rope grooves on the drum face is the rope diameter plus onesixteenth of an inch for ropes up to 2½ inches diameter. Source: Henry Broughton

• The pitch distance between rope grooves on the drum face is the rope diameter plus onesixteenth of an inch for ropes up to 1¾ inches diameter, then it increases to one-eighth of an inch. Source: Ingersoll Rand

• The pitch distance between rope grooves on the drum face may be calculated at the rope diameter plus 4% for ropes of any diameter. Source: Larry Cooper

• The maximum allowable hoop stress for drum shells is 25,000 psi; the maximum allowable bending stress for drum shells is 15,000 psi. Source: Julius Butty

• The flanges on hoist drums must project either twice the rope diameter or 2 inches (whichever is greater) beyond the last layer of rope. Source: Construction Safety Association of Ontario

Shafts and Gearing

• At installation, the allowable out of level tolerance for the main shaft of a drum hoist is onethousandths of an inch per foot of length. Source: Gary Wilmott

• Square keys are recommended for shafts up to 165 mm (6½ inches) diameter. Rectangular keys are recommended for larger shafts. Standard taper on taper keys is 1:100 (1/8 inch per foot). Source: Hamilton’s Gear book

• The width of a key should be ¼ the shaft diameter. Source: Jack de la Vergne

• For geared drives, pinion gears should have a minimum number of 12 teeth and preferably not less than 17. If the pinion has less than 17 teeth, undercutting may occur and the teeth should be cut long addendum (“addendum” is the distance between the pitch line and the crown of the tooth). Source: Hamilton’s Gear book

• For geared drive drum hoists, pinion gears should have a minimum number of 14 teeth. Source: Ingersoll Rand

Overwind and Underwind

• The overwind distance required for a drum hoist is one foot for every hundred fpm of hoist line speed. Source: Tad Barton

• The overwind distance required for a high-speed drum hoist is 7m. Source: Peter Collins

• The underwind distance required is normally equal to ½ the overwind distance. Source: Jack de la Vergne

Hoist Inertia

• The residual inertia of a double-drum hoist (including the head sheaves and motor drive, but not ropes and conveyances), reduced to rope centre, is approximately equal to the weight of 10,300m (33,800 feet) of the hoist rope. For example, the approximate inertia (WR2) of a 10-foot double drum hoist designed for 1½ inch diameter stranded ropes weighing 4 Lbs. per foot, will be:

5 x 5 x 4 x 33,800 = 3,380,000 Lbs-feet2.

Source: Tom Harvey

• The inertia of a single-drum hoist may be assumed to be 2/3 that of a double drum hoist of the same diameter. Source: Ingersoll-Rand

• The inertia (in Lbs-feet2) of the rotor of a direct current (DC) geared drive hoist motor is approximately equal to 1,800 times the horsepower of the motor divided by its speed (RPM) to the power of 1.5:

WR2 = 1800 [HP/RPM] 1.5

Source: Khoa Mai

• The inertia (in Lbs-feet2) of the rotor of a direct current (DC) direct drive hoist motor is approximately equal to 850 times the horsepower of the motor divided by its speed (RPM) to the power of 1.35:

WR2 = 850 [HP/RPM] 1.35 .

Root Mean Square Power

• Power consumption (energy portion of utility billing) of a drum hoist is approximately 75% of root mean square (RMS) power equivalent. Source: Unknown

• In calculating the RMS horsepower requirements of a drum hoist, it is not important to determine a precise value for the inertia. A 10% error in inertia results in a 2% error in the RMS horsepower. Source: Tom Harvey

Peak Power

• For a DC hoist motor, the peak power should not exceed 2.1 times the RMS power for good commutation. Source: Tom Harvey

• A typical AC induction hoist motor is supplied with a 250% breakdown torque. In application, this means that the peak horsepower should not exceed 1.8 times the RMS power. Source: Larry Gill

Delivery

• The delivery time for a new drum hoist is approximately 1 month per foot of diameter (i.e. for a 12-foot double drum hoist, the delivery time is approximately 12 months). Source: Dick Roach

• The delivery time for new wire ropes for mine hoists is approximately 4 months for typical requirements. For special ropes manufactured overseas, delivery is near 6 months. Source: Khoa Mai

3. Tricks of the Trade

• The easy way to design a drum hoist is to first determine the required hoisting speed and payload, then determine the rope that is needed to meet the safety factor (SF). The hoist parameters can then all be determined only considering the hoist rope and line speed. Source: Tom Harvey

• For purposes of initial design, the hoist line speed should be 40% of the highest speed that is theoretically obtainable over the hoisting distance. This value leaves room to increase the speed at some future date to as high as 60% without seriously compromising power costs. Source: ASEA

• The statutory minimum drum diameter to rope diameter ratios have been deleted from MSHA regulations; however, the ratios remain intact in the ANSI guidelines and these should be incorporated into the specifications for a proposed drum hoist installation at normal hoisting speeds. Source: Julian Fisher

• Where guidelines indicate an 80:1 drum to rope ratio, it may be reduced to 72:1 at hoisting speeds up to 2,000 fpm (10m/s) without significant loss of rope life when employing stranded wire ropes on drum hoists. For speeds exceeding 3,000 fpm (15m/s), the minimum drum diameter to rope diameter ratio is 96:1. At this minimum, the head sheave diameter to rope diameter ratio may be increased to 120:1 as an inexpensive means to help maintain good rope life. Source: Largo Albert

• The overwind distance is normally first calculated for the minimum statutory requirement and then increased if required to meet good engineering practice. Source: Jack de la Vergne

• For deep shafts, the overwind distance calculated must include an allowance for less turns of the hoist drum that result from hoisting an empty conveyance. Hoist controllers don’t know where the conveyance is; they precisely track the revolutions of the hoist drum. Source: Largo Albert

• The inertia of the drive motor rotor must be multiplied by the square of the gear ratio for the effect at drum radius. Source: John Maude

• An easy way to obtain an accurate value for the RMS horsepower of a counterweight hoisting system (round trip) from a computer program designed for balanced skip hoisting (one-way trip) is by making two runs. The first run hoists the full payload and the second hoists the counterweight while lowering the empty conveyance. The RMS horsepower for the round trip may then be obtained from averaging the heating values:

RMS HP = [(HP12 + HP22)/2] ½

Source: Jack de la Vergne

• An easy way to obtain a value for the RMS horsepower of a double-drum sinking hoist from a computer program designed for balanced skip hoisting is to substitute the sums of the stop and creep times in the sinking cycle for those of the skipping cycle. Source: Jack de la Vergne

• The RMS horsepower calculation is not always the criteria for selecting the drive for a drum hoist installation. When hoisting single from a deep horizon (or balanced hoisting from great depths), if the peak horsepower exceeds the RMS by a wide margin, the peak horsepower may be the basis for selecting the size of the hoist drive. Source: Jack de la Vergne

• For drum hoists, fleet angles of 1 in 45 (10 16’) or 1 in 50 (10 9’) are desirable. Source: Henry Broughton

• The fleet angle for drum hoists should not exceed 10 30’. Source: Ingersoll Rand

• In mine-shaft hoisting, the fleet angle should be as close as possible to 10 20’. Excessive drum wear and poor spooling will result if this angle is exceeded. Source: Wire Rope Industries

• Ideally, the fleet angle should not exceed 10 15’. Some line scrubbing will occur in the zone between this angle and 10 30’, but at a wider angle the rope may pull away from the flange or jump at high speed. Source: Lebus International

• In practice, the maximum acceptable fleet angle depends upon the rope line speed. For speeds of 5m/s (1,000 fpm) 10 45’ may be satisfactory as may 10 30’ at 10m/s (2,000 fpm) and 10 15’ at 15m/s (3,000 fpm). Source: Jack de la Vergne

• A minimum fleet angle of 30’ for a drum hoist will ensure that the rope will cross back and start a new layer without piling. Source: Fred Edwards

• Ideally, the minimum fleet angle should not be less than 15’. In the zone between this angle and zero, there may be trouble to kick or turn the rope back for the next layer. Source: Lebus International

• You have only to consider a tower mounted auxiliary drum hoist over a deep shaft (that operates without an intermediary sheave) to realize that the minimum acceptable fleet angle is zero. Source: Cass Atkinson

• In practice, the minimum fleet angle must never be negative, but it may be near to zero at slow hoisting speeds. Source: Jack de la Vergne

• Optimization of fleet angle geometry is obtained when the axes of the head sheaves are aligned to aim the sheave flight at the center of the face of the hoist drums rather than have the flights exactly parallel. Source: Largo Albert

4. Hoist Cycle Time “T”

One of the important aspects of hoisting is determining the cycle time. For an existing installation, it may be measured with a stopwatch or determined with a portable hoist trip recorder. The cycle time must be determined to design and specify a proposed hoist and, for this purpose, a simulated hoist cycle is calculated. The simulated cycle enables prediction of the hoist production and the capacity of the drive motor(s).

The hoist cycle time is the time taken for one complete trip. It is usually measured in seconds. The cycle is different for skipping, caging, or shaft sinking. For balanced hoisting (i.e. two skips), it is a one-way trip. For single hoisting or counterweight hoisting, the cycle is a round trip (up and down).

The cycle consists of the following components.

• Creep speed – typically 2 feet/second, except for cage hoists that creep at 1 foot/second.

• Acceleration – rate typically varies with line speed.

• Full speed – maximum rated or controlled line speed of hoist.

• Retardation – rate typically varies with line speed.

• Rest – Stop: 10-15 seconds for skip, 30-45 seconds for cage.

For hoisting skips or sinking buckets in balance, the cycle time, T (in seconds), can be accurately simplified to the following formula.

T = H/V + V/a + stops + creep times

In which

− H is the hoisting distance in meters (or feet)

− V is the full line speed in meters/second (or feet per second)

− a is the average of acceleration and retard rates in m/s/s (or feet per second/second)

Stops are rest periods at the pocket, dump, or hanging mark (in seconds), and creep times are the sum of the duration of travel at creep speed (in seconds).

The acceleration and retard rates may be adjusted for a particular installation. For purposes of general cycle calculations, it can be assumed that they are equal and proportional to the hoist speed.

a = V/15

In which

− a is feet/second/second (or m/s/s)

− V is feet per second (or m/s)

This permits a further simplification that is satisfactory to determine the hoist cycle for balanced hoisting.

T = H/V + 15 + stops + creep times 

Stops

The stops for balanced hoisting with skips include simultaneous loading and dumping. This is traditionally assumed to be 10 or 12 seconds but ought to be increased to 15 seconds or more when automatic hoisting is employed. The extra time is required for PLC proving before and after the skip is loaded. The stop time for cage hoisting is taken at 30 seconds for a small cage and 45 seconds for a large one. The sum of the stop times for double-drum shaft sinking may be taken as 45 seconds.

Creep Times

The creep times for skip hoisting applications is usually taken as equal to 5 seconds at the beginning and 5 seconds at the end of the wind (“creep out” and “creep in”). For deep shafts, the creep out can be omitted, but the creep in is typically increased to 15 or even 20 seconds for high-speed hoisting from deep shafts, to provide an extra safety margin. For cage hoisting, the creep out can be omitted, but the creep in may be increased to 10 seconds to allow for spotting the deck. The sum of the creep in times for shaft sinking in North America with a double-drum hoist may be taken as 65 seconds, and for creep out it totals about 40 seconds.

With these considerations, formulas for the hoisting cycles (in seconds) of different drum hoisting applications (valid for use with either metric or Imperial units) can be derived as follows. 

Typical shaft skip hoisting in balance

T = H/V + 35 (manual)

Typical shaft skip hoisting in balance

T = H/V + 40 (automatic)

Deep shaft skip hoisting in balance

T = H/V + 45 (automatic)

Shaft sinking in balance

T = H/V + 165 (North America)

Shaft sinking in balance

T = H/V + 135 (South Africa*)

Small cage and counterweight hoisting

T = 2H/V + 100 (round trip)

Large cage and counterweight hoisting

T = 2H/V + 130 (round trip)

Single drum shaft sinking (North America)

T = 2H/V + 215 (round trip)

* South African shaft sinkers employ a creep speed higher than 2 feet/second.

Note

At installations where skips are hoisted on rope guides, the cycle time may have to be modified to account for slow down at the ends of travel required for the transition from rope guides to fixed guides. An entry speed of 300 fpm (1.5m/s) is considered desirable although there are installations that have been carefully engineered to permit a faster transition speed (as high as 1,100 fpm).

5. Maximum Line Speeds for Drum Hoists

The speed selected for a proposed drum hoist installation is first selected on the basis of economics. This speed is determined with sufficient accuracy by applying a rule of thumb formula (Cooper or Ingersoll Rand formulas provided above). A practical limit exists to the hoisting speed that may be employed. This maximum speed may be determined by rule of thumb and by investigating the maximum speeds employed at existing operations elsewhere.

Case Histories

Tables 13-1 and 13-2 show case histories on hoisting speeds on wood and fixed guides.

Table 13-1 Hoisting Speeds on Wood Guides – Exceeding 10m/s (2,000 fpm)

Table 13-2 Hoisting Speeds on Fixed Guides (Steel), Exceeding 15m/s (3,000 fpm)

6. Production Availability

Confusion and controversy exists in the mining industry when defining the word “availability” as applied to mine hoists. For hoist maintenance personnel, it may mean the percent of the time the piece of equipment is available to work compared with the total time available. On the other hand, those engaged in selecting and evaluating hoists for mine service must consider the availability of the total hoist system, taking not only maintenance downtime into account, but also downtime due to shaft repairs, power outages, rope dressing, skip change-out, etc. This chapter is concerned with the availability of the total system, and for this purpose, it is described as “production availability.”

To determine the production availability of a double drum hoist for purpose of estimating hoisting capacity per day, a detailed calculation should be made for each case, taking into account the total hoisting system. This will include allowances for empty loading pocket, full bin, hoisting spill, etc. It will usually be equivalent to approximately 16 hours of hoisting per day (67%), for a seven day per week operation. For a five or six day per week operation, it may be 18 hours per day (75%) because some maintenance work can be performed on the weekend.

Example

Facts: 

1. Estimate is based on a seven-day workweek

2. Automatic hoisting is assumed.

3. No cage service is required

4. 12-day annual shutdown is assumed

Solution:

Hoist plant availability is shown in Table 13-3.

Table 13-3 Hoist Plant Availability – Double Drum Hoist (Seven Days per Week Operation)

Example

Determine the skip capacity required for a double-drum hoist.

Facts: 

1. Production Availability = 65%

2. Production capacity required = 4,500 tpd (ore and rock)

3. Hoisting distance (lift), H = 2,025 feet

4. Fully automatic hoisting

5. 24 hours per day operation

Solution: 

1. Optimum line speed, V = 44 x √2025 =1,980 fpm = 33FPS

2. Cycle Time, T = H/V + 40 = 2025/33 + 40 ≈ 100 seconds

3. Trips per hour = 3,600/100 = 36

4. Trips per day = 36 x 24 x 65% = 562

5. Skip capacity = 4,500/562 = 8 tons

Note

Refer to Chapter 15 of this handbook to determine the hoist rope required for a skip capacity of 8 tons and subsequent determination of drum hoist design parameters for this example.

Drum Hoists

1. Introduction

Drum hoists are employed in mines on tuggers, slushers, stage winches, cranes, rope tensioners, and even for long plumb line winches. Chapter 13 is mainly devoted to drum hoists that serve as mine hoists (winders). These machines are the most significant hoists in a mine, used for hoisting the ore and waste rock as well as moving personnel, equipment, and materials into and out of the mine.

Single-drum mine hoists are satisfactory for limited application; however, most are manufactured double drum to facilitate balanced hoisting of two conveyances in the shaft. Balanced hoisting can be accomplished with a single-drum hoist for shallow applications that require a single layer of rope on the drum. In this case, the rope being wound is wrapped in the same grooves that are vacated by the other rope being unwound. Single-drum hoists have been built with a divider flange and even with the drums of different diameter on either side of the flange (“split-differential”) to accomplish balanced hoisting – these designs are no longer manufactured. This chapter is largely devoted to the double-drum mine hoist because it is by far the most common type employed.

All mine hoists manufactured today are driven electrically by motors that have an independent ventilation source. Having an independent source reduces the horsepower requirements by more efficient cooling of the windings especially during slow-speed operations and permits filtering of the air that reaches the motor.

Until recently, DC drives with solid-state converters (thyristors) were almost exclusively employed. Lately, larger mine hoists have been manufactured with AC drives that are frequency controlled (cyclo-converter). Typically, the larger double drum hoists are direct driven with overhung armatures, while double helical open gears drive those of medium size. Small or very slow mine hoists may employ a gearbox for speed reduction.

For mine applications, drum hoists compete with friction hoists (refer to Chapter 14). The decision concerning which one is best employed for a particular application is discussed as an example of a side study in Chapter 6 – Feasibility Studies. Some hoisting parameters explained in this chapter (e.g. hoist cycle time) have equal application to friction hoists.

For historical reasons, drum hoists (unlike friction hoists) are still thought of in terms of Imperial rather than metric units. To describe the size of a drum hoist, miners will say “a 10-foot hoist” rather than “a 3m hoist.” For this reason, the explanations and calculations that follow are mainly performed in Imperial units. The Blair multi-rope (BMR) hoist (a variation of the double-drum hoist) employed for extremely deep shafts is not discussed in this Chapter.

2. Rules of Thumb

Hoist Speed

• The maximum desirable speed for a double-drum hoist with fixed steel guides in the shaft is 18m/s (3,600 fpm). Source: Peter Collins

• The maximum desirable speed for a drum hoist with wood guides in the shaft is 12m/s (2,400 fpm). Source: Don Purdie

• An analysis of the theory developed by ASEA (now ABB) leads to the conclusion that the optimum speed is a direct function of the square root of the hoisting distance. Applying the guideline of 50% and assuming reasonable values for acceleration and retardation leads to the following rule of thumb equation for the optimum economic speed for drum hoists, in which H is the hoisting distance.

Optimum Speed (fpm) = 44H½ , where H is in feet

Or, Optimum Speed (m/s) = 0.405 H½ , where H is in metres

Source: Larry Cooper

• Assuming reasonable values for acceleration gives the following rule of thumb equations for the design speed of drum hoists, in which H is the hoisting distance (feet).

Design Speed (fpm) = 34 H½ , hoisting distance less than 1,500 feet

Design Speed (fpm) = 47 H½ , hoisting distance more than 1,500 feet

Source: Ingersoll-Rand

• The hoist wheel rotation at full speed should not exceed 75 revolutions per minute (RPM) for a geared drive, nor 100-RPM for a direct drive. Source: Ingersoll-Rand

Hoist Availability

• With proper maintenance planning, a drum hoist should be available 19 hours per day for a surface installation, 18 for an internal shaft (winze). Source: Alex Cameron

• A drum hoist is available for production for 120 hours per week. This assumes the hoist is manned 24 hours per day, 7 days per week, and that muck is available for hoisting. Source: Jack Morris

• The total operating time scheduled during planning stages should not exceed 70% of the total operating time available, that is 16.8 hours per day of twenty-four hours. Source: Tom Harvey

• In certain exceptionally well-organized shafts, utilization factors as high as 92% have been reported, but a more reasonable figure of 70% should be adopted. With multi-purpose (skipping and caging) hoists, the availability will be much lower. Source: Fred Edwards

Rope Pull

• The manufacturer’s certified rope pull rating for a drum hoist assumes the rope flight angle is 25 degrees or more from the horizontal. The rope pull rating should be reduced by 10% for an installation where the ropes run horizontally between the hoist and the head sheave. Source: Ingersoll-Rand

Hoist Drums

• The hoist drum should be designed to coil rope for the hoisting distance plus an allowance equal in length to 10 dead wraps on the drum. Source: John Stephenson

• The hoist drum should be designed to coil sufficient rope for the hoisting distance plus an allowance of 500 feet, for most applications. Very deep shafts may need 600 feet of allowance. Source: Jack de la Vergne

• The hoist drum should be designed to coil sufficient rope for the hoisting distance plus the statutory three dead wraps, the allowance for rope cuts and drum pull-ins for the life of the ropes plus at least 200 feet of spare rope. (At least 250 feet of spare rope is desirable for deep shafts.) Source: Largo Albert

• The pitch distance between rope grooves on the drum face is the rope diameter plus onesixteenth of an inch for ropes up to 2½ inches diameter. Source: Henry Broughton

• The pitch distance between rope grooves on the drum face is the rope diameter plus onesixteenth of an inch for ropes up to 1¾ inches diameter, then it increases to one-eighth of an inch. Source: Ingersoll Rand

• The pitch distance between rope grooves on the drum face may be calculated at the rope diameter plus 4% for ropes of any diameter. Source: Larry Cooper

• The maximum allowable hoop stress for drum shells is 25,000 psi; the maximum allowable bending stress for drum shells is 15,000 psi. Source: Julius Butty

• The flanges on hoist drums must project either twice the rope diameter or 2 inches (whichever is greater) beyond the last layer of rope. Source: Construction Safety Association of Ontario

Shafts and Gearing

• At installation, the allowable out of level tolerance for the main shaft of a drum hoist is onethousandths of an inch per foot of length. Source: Gary Wilmott

• Square keys are recommended for shafts up to 165 mm (6½ inches) diameter. Rectangular keys are recommended for larger shafts. Standard taper on taper keys is 1:100 (1/8 inch per foot). Source: Hamilton’s Gear book

• The width of a key should be ¼ the shaft diameter. Source: Jack de la Vergne

• For geared drives, pinion gears should have a minimum number of 12 teeth and preferably not less than 17. If the pinion has less than 17 teeth, undercutting may occur and the teeth should be cut long addendum (“addendum” is the distance between the pitch line and the crown of the tooth). Source: Hamilton’s Gear book

• For geared drive drum hoists, pinion gears should have a minimum number of 14 teeth. Source: Ingersoll Rand

Overwind and Underwind

• The overwind distance required for a drum hoist is one foot for every hundred fpm of hoist line speed. Source: Tad Barton

• The overwind distance required for a high-speed drum hoist is 7m. Source: Peter Collins

• The underwind distance required is normally equal to ½ the overwind distance. Source: Jack de la Vergne

Hoist Inertia

• The residual inertia of a double-drum hoist (including the head sheaves and motor drive, but not ropes and conveyances), reduced to rope centre, is approximately equal to the weight of 10,300m (33,800 feet) of the hoist rope. For example, the approximate inertia (WR2) of a 10-foot double drum hoist designed for 1½ inch diameter stranded ropes weighing 4 Lbs. per foot, will be:

5 x 5 x 4 x 33,800 = 3,380,000 Lbs-feet2.

Source: Tom Harvey

• The inertia of a single-drum hoist may be assumed to be 2/3 that of a double drum hoist of the same diameter. Source: Ingersoll-Rand

• The inertia (in Lbs-feet2) of the rotor of a direct current (DC) geared drive hoist motor is approximately equal to 1,800 times the horsepower of the motor divided by its speed (RPM) to the power of 1.5:

WR2 = 1800 [HP/RPM] 1.5

Source: Khoa Mai

• The inertia (in Lbs-feet2) of the rotor of a direct current (DC) direct drive hoist motor is approximately equal to 850 times the horsepower of the motor divided by its speed (RPM) to the power of 1.35:

WR2 = 850 [HP/RPM] 1.35 .

Root Mean Square Power

• Power consumption (energy portion of utility billing) of a drum hoist is approximately 75% of root mean square (RMS) power equivalent. Source: Unknown

• In calculating the RMS horsepower requirements of a drum hoist, it is not important to determine a precise value for the inertia. A 10% error in inertia results in a 2% error in the RMS horsepower. Source: Tom Harvey

Peak Power

• For a DC hoist motor, the peak power should not exceed 2.1 times the RMS power for good commutation. Source: Tom Harvey

• A typical AC induction hoist motor is supplied with a 250% breakdown torque. In application, this means that the peak horsepower should not exceed 1.8 times the RMS power. Source: Larry Gill

Delivery

• The delivery time for a new drum hoist is approximately 1 month per foot of diameter (i.e. for a 12-foot double drum hoist, the delivery time is approximately 12 months). Source: Dick Roach

• The delivery time for new wire ropes for mine hoists is approximately 4 months for typical requirements. For special ropes manufactured overseas, delivery is near 6 months. Source: Khoa Mai

3. Tricks of the Trade

• The easy way to design a drum hoist is to first determine the required hoisting speed and payload, then determine the rope that is needed to meet the safety factor (SF). The hoist parameters can then all be determined only considering the hoist rope and line speed. Source: Tom Harvey

• For purposes of initial design, the hoist line speed should be 40% of the highest speed that is theoretically obtainable over the hoisting distance. This value leaves room to increase the speed at some future date to as high as 60% without seriously compromising power costs. Source: ASEA

• The statutory minimum drum diameter to rope diameter ratios have been deleted from MSHA regulations; however, the ratios remain intact in the ANSI guidelines and these should be incorporated into the specifications for a proposed drum hoist installation at normal hoisting speeds. Source: Julian Fisher

• Where guidelines indicate an 80:1 drum to rope ratio, it may be reduced to 72:1 at hoisting speeds up to 2,000 fpm (10m/s) without significant loss of rope life when employing stranded wire ropes on drum hoists. For speeds exceeding 3,000 fpm (15m/s), the minimum drum diameter to rope diameter ratio is 96:1. At this minimum, the head sheave diameter to rope diameter ratio may be increased to 120:1 as an inexpensive means to help maintain good rope life. Source: Largo Albert

• The overwind distance is normally first calculated for the minimum statutory requirement and then increased if required to meet good engineering practice. Source: Jack de la Vergne

• For deep shafts, the overwind distance calculated must include an allowance for less turns of the hoist drum that result from hoisting an empty conveyance. Hoist controllers don’t know where the conveyance is; they precisely track the revolutions of the hoist drum. Source: Largo Albert

• The inertia of the drive motor rotor must be multiplied by the square of the gear ratio for the effect at drum radius. Source: John Maude

• An easy way to obtain an accurate value for the RMS horsepower of a counterweight hoisting system (round trip) from a computer program designed for balanced skip hoisting (one-way trip) is by making two runs. The first run hoists the full payload and the second hoists the counterweight while lowering the empty conveyance. The RMS horsepower for the round trip may then be obtained from averaging the heating values:

RMS HP = [(HP12 + HP22)/2] ½

Source: Jack de la Vergne

• An easy way to obtain a value for the RMS horsepower of a double-drum sinking hoist from a computer program designed for balanced skip hoisting is to substitute the sums of the stop and creep times in the sinking cycle for those of the skipping cycle. Source: Jack de la Vergne

• The RMS horsepower calculation is not always the criteria for selecting the drive for a drum hoist installation. When hoisting single from a deep horizon (or balanced hoisting from great depths), if the peak horsepower exceeds the RMS by a wide margin, the peak horsepower may be the basis for selecting the size of the hoist drive. Source: Jack de la Vergne

• For drum hoists, fleet angles of 1 in 45 (10 16’) or 1 in 50 (10 9’) are desirable. Source: Henry Broughton

• The fleet angle for drum hoists should not exceed 10 30’. Source: Ingersoll Rand

• In mine-shaft hoisting, the fleet angle should be as close as possible to 10 20’. Excessive drum wear and poor spooling will result if this angle is exceeded. Source: Wire Rope Industries

• Ideally, the fleet angle should not exceed 10 15’. Some line scrubbing will occur in the zone between this angle and 10 30’, but at a wider angle the rope may pull away from the flange or jump at high speed. Source: Lebus International

• In practice, the maximum acceptable fleet angle depends upon the rope line speed. For speeds of 5m/s (1,000 fpm) 10 45’ may be satisfactory as may 10 30’ at 10m/s (2,000 fpm) and 10 15’ at 15m/s (3,000 fpm). Source: Jack de la Vergne

• A minimum fleet angle of 30’ for a drum hoist will ensure that the rope will cross back and start a new layer without piling. Source: Fred Edwards

• Ideally, the minimum fleet angle should not be less than 15’. In the zone between this angle and zero, there may be trouble to kick or turn the rope back for the next layer. Source: Lebus International

• You have only to consider a tower mounted auxiliary drum hoist over a deep shaft (that operates without an intermediary sheave) to realize that the minimum acceptable fleet angle is zero. Source: Cass Atkinson

• In practice, the minimum fleet angle must never be negative, but it may be near to zero at slow hoisting speeds. Source: Jack de la Vergne

• Optimization of fleet angle geometry is obtained when the axes of the head sheaves are aligned to aim the sheave flight at the center of the face of the hoist drums rather than have the flights exactly parallel. Source: Largo Albert

4. Hoist Cycle Time “T”

One of the important aspects of hoisting is determining the cycle time. For an existing installation, it may be measured with a stopwatch or determined with a portable hoist trip recorder. The cycle time must be determined to design and specify a proposed hoist and, for this purpose, a simulated hoist cycle is calculated. The simulated cycle enables prediction of the hoist production and the capacity of the drive motor(s).

The hoist cycle time is the time taken for one complete trip. It is usually measured in seconds. The cycle is different for skipping, caging, or shaft sinking. For balanced hoisting (i.e. two skips), it is a one-way trip. For single hoisting or counterweight hoisting, the cycle is a round trip (up and down).

The cycle consists of the following components.

• Creep speed – typically 2 feet/second, except for cage hoists that creep at 1 foot/second.

• Acceleration – rate typically varies with line speed.

• Full speed – maximum rated or controlled line speed of hoist.

• Retardation – rate typically varies with line speed.

• Rest – Stop: 10-15 seconds for skip, 30-45 seconds for cage.

For hoisting skips or sinking buckets in balance, the cycle time, T (in seconds), can be accurately simplified to the following formula.

T = H/V + V/a + stops + creep times

In which

− H is the hoisting distance in meters (or feet)

− V is the full line speed in meters/second (or feet per second)

− a is the average of acceleration and retard rates in m/s/s (or feet per second/second)

Stops are rest periods at the pocket, dump, or hanging mark (in seconds), and creep times are the sum of the duration of travel at creep speed (in seconds).

The acceleration and retard rates may be adjusted for a particular installation. For purposes of general cycle calculations, it can be assumed that they are equal and proportional to the hoist speed.

a = V/15

In which

− a is feet/second/second (or m/s/s)

− V is feet per second (or m/s)

This permits a further simplification that is satisfactory to determine the hoist cycle for balanced hoisting.

T = H/V + 15 + stops + creep times 

Stops

The stops for balanced hoisting with skips include simultaneous loading and dumping. This is traditionally assumed to be 10 or 12 seconds but ought to be increased to 15 seconds or more when automatic hoisting is employed. The extra time is required for PLC proving before and after the skip is loaded. The stop time for cage hoisting is taken at 30 seconds for a small cage and 45 seconds for a large one. The sum of the stop times for double-drum shaft sinking may be taken as 45 seconds.

Creep Times

The creep times for skip hoisting applications is usually taken as equal to 5 seconds at the beginning and 5 seconds at the end of the wind (“creep out” and “creep in”). For deep shafts, the creep out can be omitted, but the creep in is typically increased to 15 or even 20 seconds for high-speed hoisting from deep shafts, to provide an extra safety margin. For cage hoisting, the creep out can be omitted, but the creep in may be increased to 10 seconds to allow for spotting the deck. The sum of the creep in times for shaft sinking in North America with a double-drum hoist may be taken as 65 seconds, and for creep out it totals about 40 seconds.

With these considerations, formulas for the hoisting cycles (in seconds) of different drum hoisting applications (valid for use with either metric or Imperial units) can be derived as follows. 

Typical shaft skip hoisting in balance

T = H/V + 35 (manual)

Typical shaft skip hoisting in balance

T = H/V + 40 (automatic)

Deep shaft skip hoisting in balance

T = H/V + 45 (automatic)

Shaft sinking in balance

T = H/V + 165 (North America)

Shaft sinking in balance

T = H/V + 135 (South Africa*)

Small cage and counterweight hoisting

T = 2H/V + 100 (round trip)

Large cage and counterweight hoisting

T = 2H/V + 130 (round trip)

Single drum shaft sinking (North America)

T = 2H/V + 215 (round trip)

* South African shaft sinkers employ a creep speed higher than 2 feet/second.

Note

At installations where skips are hoisted on rope guides, the cycle time may have to be modified to account for slow down at the ends of travel required for the transition from rope guides to fixed guides. An entry speed of 300 fpm (1.5m/s) is considered desirable although there are installations that have been carefully engineered to permit a faster transition speed (as high as 1,100 fpm).

5. Maximum Line Speeds for Drum Hoists

The speed selected for a proposed drum hoist installation is first selected on the basis of economics. This speed is determined with sufficient accuracy by applying a rule of thumb formula (Cooper or Ingersoll Rand formulas provided above). A practical limit exists to the hoisting speed that may be employed. This maximum speed may be determined by rule of thumb and by investigating the maximum speeds employed at existing operations elsewhere.

Case Histories

Tables 13-1 and 13-2 show case histories on hoisting speeds on wood and fixed guides.

Table 13-1 Hoisting Speeds on Wood Guides – Exceeding 10m/s (2,000 fpm)

Table 13-2 Hoisting Speeds on Fixed Guides (Steel), Exceeding 15m/s (3,000 fpm)

6. Production Availability

Confusion and controversy exists in the mining industry when defining the word “availability” as applied to mine hoists. For hoist maintenance personnel, it may mean the percent of the time the piece of equipment is available to work compared with the total time available. On the other hand, those engaged in selecting and evaluating hoists for mine service must consider the availability of the total hoist system, taking not only maintenance downtime into account, but also downtime due to shaft repairs, power outages, rope dressing, skip change-out, etc. This chapter is concerned with the availability of the total system, and for this purpose, it is described as “production availability.”

To determine the production availability of a double drum hoist for purpose of estimating hoisting capacity per day, a detailed calculation should be made for each case, taking into account the total hoisting system. This will include allowances for empty loading pocket, full bin, hoisting spill, etc. It will usually be equivalent to approximately 16 hours of hoisting per day (67%), for a seven day per week operation. For a five or six day per week operation, it may be 18 hours per day (75%) because some maintenance work can be performed on the weekend.

Example

Facts: 

1. Estimate is based on a seven-day workweek

2. Automatic hoisting is assumed.

3. No cage service is required

4. 12-day annual shutdown is assumed

Solution:

Hoist plant availability is shown in Table 13-3.

Table 13-3 Hoist Plant Availability – Double Drum Hoist (Seven Days per Week Operation)

Example

Determine the skip capacity required for a double-drum hoist.

Facts: 

1. Production Availability = 65%

2. Production capacity required = 4,500 tpd (ore and rock)

3. Hoisting distance (lift), H = 2,025 feet

4. Fully automatic hoisting

5. 24 hours per day operation

Solution: 

1. Optimum line speed, V = 44 x √2025 =1,980 fpm = 33FPS

2. Cycle Time, T = H/V + 40 = 2025/33 + 40 ≈ 100 seconds

3. Trips per hour = 3,600/100 = 36

4. Trips per day = 36 x 24 x 65% = 562

5. Skip capacity = 4,500/562 = 8 tons

Note

Refer to Chapter 15 of this handbook to determine the hoist rope required for a skip capacity of 8 tons and subsequent determination of drum hoist design parameters for this example.