The employment of an optimal brake for overhead crane applications yields desirable and positive results: increased safety, greater efficiency, improved uptime, and reduced maintenance. Conversely, improper brake selection can be catastrophic.

After considerable field experience, substantive consultations with subject matter experts, and extensive professional research, the author highlights a set of critical variables to aid professionals in choosing the most suitable brake for any given overhead crane application. When carefully considered and applied, these variables will enhance crane performance and provide reliable braking performance.

IndustrIal crane brakes: the fundamentals

In steel and iron-making applications, industrial cranes perform three main motions: forward movement along the X Axis which is referred to as bridge motion, lateral/left-right movement along the Y Axis which is referred to as trolley motion, and up/down movement along the Z axis which is referred to as hoist motion.

Historically, the most typical brakes found in US steel and iron-making plants are DC magnetic drum brakes known as "clapper brakes." To set the brake, force is applied to the brake drum by brake pads which are bonded or riveted to the brake shoes. Like every brake used in overhead crane applications, DC magnetic drum brakes are fail-safe: a spring releases to "set" or apply the brake when power is interrupted.

This effectively prevents loads from moving or falling during power failures. The brake is released through the DC-powered magnetic coil, which is either shunt or series-wound. The brake is set when force is applied to the brake drum, which reduces or stops shaft rotation. In this way, a ladle loaded with molten steel can be prevented from falling vertically or making unsafe bridge or trolley movements.

The Association of Iron and Steel Engineers (AISE) established standard brake drum diameters: 8", 10". 13", 16", 23", and 30 ".

Although the majority of cranes currently in use conform to AISE standards, there are some general purpose brakes that utilize other sizes such 14" and 18" drum diameters which were common were common before the standard was established.

Because Electrical Overhead Travelling (EOT) Cranes range in carrying capacity from 25 tons to over 400 tons, the type of brakes needed to safely control a given load can vary significantly, according to the following factors: understanding the numbers involved in calculating mechanical braking torque, the real versus nominal coefficient of friction, the specific type of braking application, burnishing of the linings, type of brake being used, and additional safety considerations which may require secondary emergency brakes to be installed.

UnderstandIng the numbers

One of the primary purposes of this paper is to stress the importance of accurately specifying the appropriate brake for a given application. To accomplish this objective, it is necessary to have an understanding of the variables involved in the braking process and analyze the effects of their interaction.

A critical variable involved in calculating the specifications of a given braking application is torque. Torque can be defined as, "A measurement of the propensity of a given force to cause the object upon which it acts to twist about a certain axis. The torque is simply the product of the magnitude of the Applied Force and the length of the lever arm. " For any given industrial braking application, there will be a specific torque requirement that the brake must conform to in order to meet the demands of the crane on which it is utilized. There are fundamental torque formulas critical for specifying the optimal brake for a given application. Electrical (full-motor torque) and Mechanical Braking Torque are the two principal torque values used for these calculations.

Motor Torque is an electrical formula which refers to the maximum load the motor can produce. The following is the formula for motor torque which uses the horsepower and RPM of the motor being used and is expressed in Foot Pounds (Lb X Ft) :

Torque=(HP X 5250)/RPM (1)

The mechanical torque produced by the brake must be sufficient to overcome full motor torque, and a service factor must also be considered. According to AIST Technical Report No. 6, " Brake sizes shall be as recommended by the brake manufacturer for the service, but in no case shall the summation of all brake ratings in percent of hoist full load hoisting torque at the points of brake application be less than the following:

a) 150% when only one brake is used. b) 150% when multiple brakes are used and the hoist is not used to handle hot metal; failure of any one brake shall not reduce total braking torque below 100%. c) 175% for hoists handling hot metal; failure of any one brake shall not reduce total braking torque below 125%.

For example, if two brakes are used, each must be rated 100% of the total full load hoisting torque (125% each for hot metal). If three brakes are used, each must be rated 50% (62.5% each for hot metal). If four brakes are used, each must be rated 37.5% (43.75% each for hot metal). In each of these cases, the failure of one brake does not cause the remaining braking torque to fall below the required minimum. "

As will be further delineated, it is often prudent to use a higher service factor due to variables that can potentially decrease the anticipated "nominal" torque value and thus precipitate the need for a larger "cushion" to prevent an underestimation of torque. The following is the formula for calculating Mechanical Braking Torque which is expressed in foot-pounds (ft-lbs.):

Mechanical Braking Torque=AF x 0.42(DF)/ 24 (2)

AF= Applied Force

.42= Standard (Nominal) Coefficient of Friction

D= Disc/Drum Diameter F= Face of the Caliper/Shoe

Applied Force is the first variable used in calculating mechanical braking torque, a linear force denoted in Pounds (Lb.). Force comes from the mechanism of the brake which is the compression spring plus the mechanical advantage (lever ratio). In other words, Applied Force is the product of force times distance and is calculated by multiplying the spring force times the mechanical ratio (which is determined by the brake design). Mechanical braking torque measures the maximum torque value that can be generated by the brake and it is a calculation of the rotational (torque) value that is yielded by the amount of Applied Force it has. Fundamentally, torque is a rotational force and Applied Force is a linear force. For fail-safe springapplied brakes that are used in iron and steelmaking facilities, Applied Force is the force that the shoe/pad will contact the brake drum/disc at which is applied by a spring. Therefore, mechanical braking torque is a measure of the maximum torque value that can be produced by the brake.

You will see in the mechanical braking torque formula that there are several variables involved in converting Applied Force into torque. D is a measure of the Disc or Drum Diameter, and F is a measure of the face of the caliper/brake shoe. As will be further explained, friction coefficient is a vital component of the torque calculation and specifying the brake as a whole. While 0.42 is a common friction coefficient for linings in the context of industrial braking applications, friction linings can be furnished in coefficients that are much higher and much lower. The higher the coefficient of friction then the more you are converting Applied Force into braking force. Hypothetically, if you could convert 100% of Applied Force into braking force then you would achieve a 1.00 coefficient of friction.

Friction: More art than science

Choosing the best brake for a given application is more art than science and there is no algorithm to calculate the precise coefficients. Too often, the assumption is made that brake size is the preeminent variable in determining the amount of torque yields. It is true that the size of a brake is proportionate to the amount of torque that will be generated: the larger the brake, the greater the torque. However, there are several other variables that are extremely important in determining the torque value of a brake, and brake size is not necessarily the paramount determinant for braking torque.

When first specifying a brake for an application, there will be an ideal torque value that, ceteris paribus, will be achieved. This will be referred to as the "nominal" torque value.

However, torque is never a precise science and knowing the "real" torque value provided in an application could be considered a "subjective equilibrium." There are many different factors to be considered and achieving a desired torque value is sometimes elusive and subject to many different factors. Of these factors, especially important but often underestimated is the coefficient of the friction linings/pads. As previously stated, 0.42 is an estimated standard value that is commonly used for braking torque formulas. The coefficient of friction between two surfaces in contact is equal to the force required to overcome the friction divided by the reaction force between the two surfaces . The formula used to calculate the coefficient of friction is:

µ = F /R (3)

µ= Coefficient of Friction

F= Force Required to Overcome the Friction

R= Reaction force between the two surfaces

While friction coefficient value of 0.42 is a relatively good estimate for current metallic impregnated material, this information does not exist in published standards. Neither AIST, NEMA, nor ANSI standards contain any reference to friction coefficient or a definition of the term burnishing, which is another term that relates to friction coefficient and is prevalently used in this paper. The 0.42 coefficient of friction value is an ideal standard which could be considered the "nominal" friction coefficient. This nominal friction coefficient makes a number of assumptions, including but not limited to the brake being used in a dynamic application, the brake being operated in clean environmental conditions, and whether or not the friction linings are adequately burnished, among others. It is possible that this "real" friction value will be achieved in an application. However, to get a more realistic and accurate prediction of the friction value that will be achieved, a more intricate and strategic approach must be taken in which we calculate the "real" coefficient of friction is the true value that is achieved once all factors are considered based upon the coefficient of friction of the linings actually being used among other factors.

This number is subject to fluctuations. For instance, in an industrial environment generating a substantial amount of dirt and grime, the friction linings are likely to become contaminated and the friction coefficient will be effectively reduced. For example, friction linings that are 0.42 in a clean environment could become 0.35 after contamination. Consequently, in an application that is subject to high amounts of contamination, it may be wise to increase the service factor to compensate for the environmentally-reduced friction values.

For drum brakes, the material used in manufacturing the brake drum can also affect the "real" coefficient of friction. With the use of a ductile iron wheel or qt100 disc, the friction coefficient will not be reduced. However, certain end users may have a preference for using a stainless steel brake drum and if this be the case the friction coefficient can potentially decrease as much as twenty per cent which would consequently de-rate the mechanical braking torque.

Another indispensable constituent in determining the "real" coefficient of friction is the type of brake application itself. The two principal braking applications we must consider to determine the proper friction coefficient are Dynamic vs. Static. Service brakes are primary brakes that are in continuous use whereas secondary parking brakes are used as a backup.

Both Primary and Secondary brakes can be used in either Dynamic or Static (otherwise known as "holding") applications. Dynamic Braking is the process of reducing speed of any rotating machine. In the case of EOT crane braking systems, Dynamic Braking occurs when brakes are used to stop the load, which is achieved by stopping the rotating shaft from turning. The secondary brakes are considered holding brakes since they do not stop the load, but instead they hold the load. According to the International Society of Automation, "Holding brakes’ principal use is to safely keep a load in place in situations where the power either is turned off or fails. Holding brakes are available in two common configurations: dynamic stopping and static holding. Dynamic stopping brakes are commonly used as cycling brakes, which take the wear and tear of constant on/off engagements while the shaft is rotating and still provide long life. Static holding brakes are for simple loadholding applications."

Understanding the difference between dynamic stopping and static holding applications is significant because a different brake design is required for each respective scenario, ergo using the wrong brake can face potentially negative consequences. To elaborate, the International Society of Automation says, "A static holding brake that’s incorrectly applied in a frequent cycling application will wear out and fail quickly because it isn’t designed for wearing applications. Similarly, a unit designed as a dynamic brake needs engagements at speed to maintain its full torque rating.

A dynamic brake that’s used solely as a holding brake may experience torque degradation, reducing its performance over time due to a loss of burnish." It is imperative to understand that the coefficient of friction is different in a static holding versus a dynamic stopping application.

A dynamic stopping application will generally have a higher friction coefficient than its static counterpart. In a static application, the braking starts from a stationary position; the lining will not achieve full contact on the disc/drum because there will be high points/ridges on the linings. In a dynamic stopping application, the high points will be diminished due to the more active braking application and the enhanced grip and contact with the disc/drum yields a greater coefficient of friction. As a result of this use differentiation, it is suggested that the friction coefficient for a static application is to be de-rated approximately twenty percent from its nominal value. The same application from a stopping position would yield 0.4, but would yield 0.5 in a dynamic application.

Brake Fade

In a dynamic braking application, there is the propensity for the phenomenon of brake fade to occur and this is an important concept to factor into brake selection. To elaborate on the previous paragraph, the primary difference between static and dynamic braking is that dynamic braking applications encompass a repetitive stopping of a rotating machine and thus the brake has to absorb the kinetic energy that is built up by inertial loads. As a result, the brake must transfer this kinetic energy which results in heat build-up and wear on the surfaces of the rotating components. The brakes function by converting the kinetic energy of the rotating shaft into thermal energy during deceleration which produces a substantial amount of heat which must then be transferred into the surroundings and into the air stream. Conversely, with static holding applications, all rotating components come to a rest and the brake simply holds the load. In these applications, no heat builds up and there is very little wear.

Brake/pad fade occurs when the temperature at the interface between the pad and the disc exceeds the thermal capacity of the pad. One result of brake fade is the formation of resin components on the linings commonly referred to as "glazing" which effectively reduces the integrity of the linings and thus lowers the coefficient of friction of the linings. This is typically indicated with a telltale odor and/ or smoke. Fluid boiling and vaporization is another result of brake fade which has adverse effect on friction coefficient and ultimately the ability for the brake to stop. Gas bubbles are formed when fluid boils on the friction materials and ultimately degrade the ability for the friction material to interface with the disc/drum. This process is incremental with several warning signs.

Since brake fade occurs as a result of continuous braking and the heat that ensues, decreasing the stopping/braking time is the principal means to mitigate the risk of brake fade. Stopping time is directly proportionate to the amount of heat generated during the friction occurrence. In other words, if stopping time is decreased from four seconds to two seconds, heat generation is reduced by fifty percent. Less heat means less pad/ shoe wear. One way to reduce stopping time is by increasing the torque of the brake. As discussed, the variables that can be manipulated to create variation in the torque formulate are disc diameter, friction coefficient, and Applied Force of the brake.

By using a larger disc, using friction linings/pads with a higher coefficient of friction, or using a brake with a greater amount of Applied Force will yield a higher torque value. Contrary to popular belief, increasing pad area will not increase the brake torque. However, increasing the pad area will decrease pad wear and mitigate the risk of brake fade; larger pad area means greater surface area to dissipate heat. Theoretically, the same amount of torque can be achieved by using either an extremely large pad or an extremely small one.

However, the amount of heat being generated in a high torque application would be far too great for the smaller pad to dissipate, increasing the danger that it could burn up very quickly. There are also safety features modern brakes offer which mitigate the risks of brake fade. Selfadjusting mechanisms are ideal because they compensate for lining wear and, as a result, torque does not decrease due to lining wear. Lining wear indicators are helpful, too; they provide a signal when lining thickness reaches a critical threshold and indicate that linings must be changed.

Pursuant to the concept of static versus dynamic braking, another critical component in calculating the "real" coefficient of friction value is the concept of burnishing. One definition of burnishing of brake linings holds that "burnishing is a method of conditioning the top layer of the friction compound during the manufacturing process removing the need for the bedding-in. This effect is achieved by process equipment that subjects the friction surface, in a controlled environment, to a short burst of heat exposure in the temperature range where green fade occurs. Burnishing is therefore a method of preconditioning the brake lining rendering it ready for full service upon installation.

The user is in a position to install and drive without having to follow complicated bedding-in procedures, although a certain amount of "self-bedding-in" is still present. " While some original equipment manufacturers attempt to burnish friction linings during the manufacturing process, it is virtually impossible to achieve complete burnishing until the brake is installed into its true application due to application-specific variables that cannot be accounted for at the manufacturing facility. These variables include, but are not limited to:

1. Height of the brake: if a brake is offset and not perfectly centered, the pads will not contact as designed. If the brake is too high or skewed to one end, there will be unbalanced contact and the pads will not achieve optimal contact.

2. Brake mounting: if the brake is not centered and exactly aligned, the brake drum will not be centered as the manufacturer intended; subsequently the linings will not contact the drum at the optimal angle and provide expected performance.

3. Concentricity of the brake drum (or flatness of the disc): if the brake drum is perfectly round (or if the disc is perfectly flat), then there will be no difference between burnishing at the manufacturing site versus at the live application.

However, if the drum being used is not perfectly round then the linings will contact the drum imprecisely and fail to deliver desired performance.

Once friction linings become burnished, the coefficient of friction will increase by eliminating the high/uneven points on the surface of the linings which yields greater contact and therefore greater friction. With unburnished linings, the linings will have high points. When the brake is applied these high points will contact the brake wheel but the thinner points will achieve no contact. If burnished perfectly, linings will achieve one hundred percent contact. In light of this information, it can be implied that, when first put into service, linings will yield a lower amount of torque due to having less contact with the mating surface such as the drum, potentially causing more slip. After friction linings are thoroughly burnished, the high points of the linings are eliminated; they wear off and effectively become much smoother and achieve close to 100 percent contact with the mating surface, and more contact means greater torque.

At first glance, it may seem that bigger brakes provide better braking. Actually, a smaller-diameter brake can generate as much torque as a larger one by simply increasing the coefficient of friction linings applications and/or enlarging the disc diameter the calipers mate to. The following is an empirical evaluation of friction and disc diameter:

Scenario One: (using a 100-HP Motor with 1800 RPM on the low-speed side):

– Full motor torque: (5250 X 100) / 1800 = 292

– Using Service Factor of 1.5: 292 X 1.5 = 438

– Using 10:1 Gearbox Ratio: 438 X 10 = 4,380 ft-lbs of Full Motor Torque

To determine the appropriate brake to accommodate this torque requirement, calculate the Static Torque of the brake. Using a 28" disc with 3.4" pad width (manufacturer specification) in conjunction with a brake with Applied Force of 8,520 Lb. and 0.4 coefficient of friction: 8,520 X 0.4(28-3.4 ) / 24 = 3,493.20 ft-lbs. That torque is insufficient: 3,493.20 < 4,380.00 Scenario Two: Using the same disc diameter but with a larger brake which furnishes Applied Force of 12,100 ft-lbs: 12,100 X 0.4(28-4.7) / 24 = 4,698.83 ft-lbs This brake would provide sufficient torque for the application.

Under these circumstances, the smaller brake will not suffice. However, using the identical setup as Scenario One but using friction pads with a 0.6 coefficient of friction and increasing the disc diameter to 31.5" provides an Applied Force of 8,520 Lb X 0.6(31.5-3.4) / 24= 5,985.3 ft-lbs.

This brake that is more than capable of handling this torque value. In fact, the disc diameter can be decreased to 24" and still generate 4,387.80 ft-lbs of torque, which is more than sufficient for the given application. This solution has the added advantage of requiring less clearance the 28" disc and provides a degree of cost savings, as well.

The implication of these scenarios is critical: if a torque calculation for a given brake uses the nominal coefficient of friction (0.42) without taking other variables into consideration, there is the potential for the brake to underperform. A 0.05 decrease in friction coefficient may not necessarily yield a substantive difference, but in some cases it could reduce the torque value to the point of not being capable of carrying the load and thus the results would become catastrophic. Additionally, stopping power varies as a result of the application. For instance, is it desirable to have an immediate hard stop or a gradual and smoother stop?

The standard 0.42 friction coefficient may not necessarily yield the desired results for a given application. Consider a comparison of a hoist versus trolley application. For a trolley application, the stop can potentially become too "grabby" with too much friction which causes a hard stop and can wobble the load and cause its instability. This goes back to the concept of "subjective equilibrium:" the objective for bridge and trolley application is to achieve sufficient torque to stop the brake but having too much torque is a liability. Therefore, using a medium friction coefficient is ideal to avoid a situation where the linings are too "grabby." For a hoist application a "hard "stop is not a liability, and it may actually be considered an asset since the objective is to get as much torque as possible to stop the load from falling. In these instances a higher coefficient of friction may be desirable.

Another significant implication of the potential torque fluctuation is the concept of service factor. As previously stated, there are standard service factors specified by the CMAA, OSHA, and AISE which are a percentage of full motor torque. Ostensibly, these standard service factors can be used to calculate a mechanical braking torque value that is more than sufficient to handle the load in the given application.

However, being that there are several variables that contribute to the "real" mechanical braking torque value which can be inherent or extrinsic to the braking system, these service factors may not be sufficient to provide the safest and most accurate full motor torque value for a given application. For instance, if a 1.25 service factor were used as a product of full motor torque but extrinsic variables such as environmental (i.e. high degree of contamination), and intrinsic/ application-specific variables (i.e. assuming higher coefficient of friction despite static application with lack of burnishing) are not taken into consideration, it is quite possible that the 1.25 service factor may be quickly downgraded to 1.00 or lower.

If this were the case, the risk of failure would increase because the "security blanket" that a safety factor allows would be eliminated. Due to the magnitude of service factors in the safety of an industrial braking system, it is imperative that the parties involved in specifying the braking system perform an application-specific analysis to take all factors into consideration to use the best and most realistic service factor to ensure the utmost safety and reliability