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.