Digital Magazine

Controlling Your Web from Beginning to End

In Part 1 of a two-part series, Bill Hawkins looks at the devices and designs that can help you keep your web under control.

Dynamic forces that act on the running web in your converting machine may cause that web to behave in many different ways, some of which definitely can be detrimental to the product you are trying to make.

Web control devices must be installed to correct the negative action of these dynamic forces. This two-part article will cover the important roles that a few specific control devices play to keep the web stable.

Where It Starts
Web tension control begins at the payoff roll (the roll on the unwind before any processing occurs). For best results with any converting process, web payoff tension should be kept as near constant as possible from full roll to the last wraps near the core. Most modern converting machines have either a dancer or a load cell roll that is used to sense the web payoff tension.

The dancer roll changes its position as the web tension is increased or decreased. Normally, its running position is the point where the web tension (and gravity force if not nullified by the position of the roll support arms) is balanced by the opposing force of the system's actuator(s). The running position determines the amount of reverse web tension (brake pressure or brake motor regenerative amps) to apply to the unwind roll.

An encoder on the dancer roll pivot shaft determines the amount and polarity (direction) of the correction signal to apply to the brake or motor regulator so that the dancer roll can be returned to the desired running position. When the roll is kept near the running position, the payoff tension is kept nearly constant.

The load cell roll is a sensing roll that has strain gauges (or other force-sensitive devices) installed in either the roll mountings or in the roll shaft.

Both types of these web tension sensor rolls require two other rigidly mounted rolls to define a stable thread path around the sensing roll. And both provide feedback signals that are proportional to the force applied to the sensing roll and input them to their respective tension control panel. The signals may be digital or analog, depending on the specific control scheme of the panel.

The input signal is compared to a reference signal (voltage for analog or numerical value for digital) in the control panel. The reference signal magnitude is determined by the selection of an operating tension level in the panel program. Any difference between the feedback and reference signals is sent to the output regulator. The output regulator adjusts the signal that it has been supplying to the control panel of the actuator device of the unwind brake (or unwind motor panel if the motor is being used as a brake). The brake or motor responds to keep the feedback and reference signals at the same value.

Keeping the Web Taut
I have always preferred the dancer roll to the load cell roll as a web tension sensor for the unwind process because of its ability to keep the web taut when path length changes occur due to unwind roll eccentricity or droop in loosely wound payoff rolls.

Web steering and spreader rolls are more efficient when the web is kept taut. I still recommend the dancer sensor system over the load cell roll system for high line speeds (more than 500 fpm) or whenever a passive brake (friction, magnetic particle, etc.) is used for an unwind application.

However, continued developments in rotational position controls for DC and AC flux vector motors have enabled them to reduce motor slip and keep the web taut on the guide rolls when mild eccentricity exists and the line speed is low.

For success when using a load cell sensor on the unwind stand, there must be no measurable slack in the drive train between the unwind motor and the payoff roll mandrel. And there must be additional horsepower installed to accelerate and decelerate the payoff roll in rhythm with the changes in angular velocity to keep the payoff tension near constant.

The horsepower requirement is not linear with line speed. For example, about 31/3 more horsepower is needed by the unwind motor to keep the web near constant tension when a 28-in.-dia, 1,000-lb mill roll with 0.125-in. eccentricity is paying off at 600 fpm line speed. But about 15 additional hp is needed when the line speed is increased to 1,000 fpm.

Making Dancer Rolls Perform
Two keys to good dancer performance are to keep the moving mass as low as possible while keeping the system as stiff as possible. Pivoting dancer rolls usually can fit the criterion of the design more easily than can parallel moving dancer roll assemblies mounted on siding rails.

For best performance, the pivoting support arms should be mounted on a very stiff pivot shaft. The shaft should be split between the arms and the ends connected with a zero backlash rotational aligning device. Harmonic gear mechanisms that can be locked in position are an excellent choice for aligning the two shafts.

And, the arms should not be counterweighted to balance the roll weight. Counterbalancing the dancer roll arms adds to the rotational inertia of the system, and this action opposes rotary movement of the arms and reduces the efficiency of the dancer.

Mounting the dancer vertically removes most of the gravitational force component from the forces that the system actuator must null during operation.

A good dancer roll design uses an air actuator(s) with very low friction to oppose the working forces of the web. The pivot length of the dancer roll lever arms should be about 2x the actuator arm pivot length.

In addition, the arms should be fairly long (30 in. or longer if possible), so that the actuator experiences small gravity force changes when working in the normal operating arc.

There is less angular change in the resultant force from the web as it goes from maximum to minimum tension during operation when longer arms are used. The arms should be minimum weight to reduce rotational inertia.

The air actuator(s) should have modifications made to open the exhaust (non-pressure) sides to the atmosphere so that air will enter and exit the cylinder(s) freely without increasing or decreasing the force on that side of the piston(s). The roll should be made as light and stiff as possible for the web span.

The surface should be textured to keep good friction contact with the web. Since the roll usually operates with about 180-deg wrap, the roll shape may be made concave to give web spreading at this location.

Load cell sensor rolls use much less real estate than dancer rolls. This advantage is sometimes the main reason they are incorporated into the unwind design. They also can operate with significantly less wrap angle than the dancer roll, but I would recommend that the wrap angle be not less than 90 deg.

About Load Cells
Like the dancer roll, the load cell roll can be textured and made with a concave surface for spreading the web. However, when a load cell roll sensor is used, the load cell rating must be high enough to withstand the force of a wrap—and ultimately a web break—without losing calibration. There are several load cell designs on the market now that are rugged enough to withstand the abuse of being located next to the unwind stand and still hold their calibration.

The above analysis is offered as an aid to help you choose which web tension sensor/brake system best fits your unwind stand needs and your budget. I prefer the driven unwind with either DC or AC flux vector motors to the passive brake for all light-gauge webs that are less than 1 mil thick or with either type of web tension sensor when the webs are stretched easily.

In my opinion, the higher cost of the driven unwind shaft system offsets the lower cost benefit of a frictional brake system by yielding better quality rewound rolls with less waste.

The January 2001 issue of PFFC will feature Part 2, covering web guiding from the unwind stand; steering assemblies and sensors; tension sensing at the windup; and multiple rolls on one mandrel.

The Parameters
A 1,000-lb mill roll (28-in. outer diameter (O.D.) on 6-in. core with 1/2-in. wall, line speed constant at 600 fpm, (eccentricity) true roll center of roll is displaced 0.125 in. from the turning center.

  • Revolutions/sec = ((600 x 12) /60)/28 x Pi) =1.364 RPS or SPR = 1/1.364 = 0.733 sec.
  • Using the machine as a reference point in this example, the greatest rotational velocity is when the roll eccentricity is aligned with point Y at 270 deg. This is because the turning radius at payoff point has been reduced by the amount of the roll eccentricity, and the roll rotational speed has been increased to compensate. Thus, maximum acceleration (and horsepower) must be applied between 180 and 270 deg.
  • Time for 1/4 revolution = 0.733/4 = 0.183 sec.
  • Web velocity vector at payoff point given as constant, + Vpo = (600 x 12)/60 = 120 in./sec @180 deg.
  • Rotational eccentricity, e = 0.125 in.
  • Distance that payoff point moved to the right (away from the windup) in the 1/4 revolution when the actual roll center (point D) has moved under the turning center Q, = eccentricity = 0.125 in.
  • Velocity payoff point toward the right at this moment, - Vpo = Spo/t = - 0.125/0.183 = - 0.683 in./sec. The roll tangential velocity at the payoff point must be increased by the same amount to keep payoff velocity at 120 in./sec.
  • Roll tangential velocity at the payoff point during maximum rotational velocity, + V T = Vpo + Vz = 120 + 0.683 = 120.683 in./sec.
  • Rotational velocity of payoff point at this moment, wY = 120.683/ 13.875 (roll diameter minus eccentricity) = 8.698 radians/sec.
  • Average rotational velocity between 180 and 270 deg, wAVE. = (wY + wX)/2 = (8.698 + 8.571)/2 = 8.634 radians/sec.
  • Average roll acceleration to change rotational velocity, ALPHAAVE. = (2 x (wAVE. - wX))/t = (2 x (8.634 - 8.571))/0.183 = 0.689 radians/sec ²
  • Payoff roll rotational inertia, Ic = (1/2m) X (r1² + r2²), r1 = 3.50 in., r 2 = 14.00 in., m = 1000/386 slug (in./sec² ) = 2.591 slug (in./sec²), m/2 = 1.295, r1²; = 12.250, r 2² = 196.00, Ic = 269.754 slug (in./sec²).
  • Axial torque required to produce the necessary rotational acceleration, L = Ic X ALPHAAVE. = 269.754 X 0.689 = 185.861 in #s.
  • The roll rotation speed is on average, 1.364 x 60 = 81.84 rpm.
  • Horsepower to accelerate roll in the first quadrant, hp = (185.861 x 81.84)/5252 = 2.896 hp.
Let us assume for this case that (Ic) for acceleration (L) of the core, mandrel, gears, and motor is 40 slug (in./sec2). Check your machine literature, because it may contain data for Ic of the motor and gears. When the inertia of the other parts is added to the above equations, the required unwind motor reserve hp is about 31/3 for 0.125 in. eccentricity.

Using the above equations and logic, about 15 hp is required to keep the payoff web taut when the line speed is raised to 1,000 fpm.

The above method can be used to approximate the additional horsepower needed for your unwind stand motor if you are considering using the load cell sensor system for unwinding mill rollstock.

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