To obtain definitive information on packing friction, the joint ESA/FSA Technical Task Force commissioned CETIM to carry out a follow-up project. It consists of the design and manufacture of a dedicated test rig to carry out testing in accordance with the procedure, including highly accurate systems to directly measure the frictional force of the packing alone.
The test rig is designed to test both compression packings and mechanical seals so direct comparison can be made under the same conditions (see Image 1, page 98). A torque meter is used to record the mechanical seal or packing friction on the shaft. Measurements of torque, temperature and leakage levels are recorded, and the instrumentation permits continuous monitoring of all parameters throughout the test.
After initial trials to validate the equipment functionality and accuracy of the monitoring devices, the first tests were carried out on the same graphite/ePTFE packing that had been widely used during the earlier test program.
Testing was conducted at different rotational speeds and pressures, with varying target leak rates. For direct comparison, a typical unitized, single-spring elastomer bellows mechanical seal was also tested under a range of conditions. It is an unbalanced mechanical seal with carbon graphite versus chromium oxide seal faces. The results of these tests are shown graphically in Figure 3a.
The measured torque is plotted for different water pressures, in the case of the packing with the associated shaft leak. During these tests, the gland leak rate was of the same order of magnitude as that of the shaft.
These results were unexpected. The figures for packing were much lower than predicted and were of the same order of magnitude as, and generally lower than, the mechanical seal. Of course, a degree of leakage must be tolerated when using packings, and the lubrication afforded by the leaking fluid will reduce the friction. But even when the leak rate is extremely low, as in the case at 6 bar and 1500 rpm, the friction recorded was the same as that for the mechanical seal at a lower pressure.
Rigorous checks were carried out to ensure the accuracy of the results. In particular, the measurement range of the torque meter was revised to ensure accuracy at these much lower torque levels, and it was verified that the torque levels measured for mechanical seals were generally in line with the manufacturer's published data.
A further series of tests was carried out on two other packing types and four mechanical seal variants. The packings were a lubricated natural ramie fiber, which would normally be used where higher leakage would be acceptable, and a synthetic aramid yarn packing.
The mechanical seals were one unbalanced and two balanced component seals and a cartridge balanced seal. They were chosen to represent a cross section of commonly used designs. These featured carbon graphite versus silicon carbide seal faces. This face combination is typically chosen for its low coefficient of friction. The designs had different balance ratios, and two had a composite narrow seal face and the other two had a monolithic narrow seal face.
All tests in this sequence were carried out at 6 bar pressure. The comparative results are shown in Figure 3b.
Some of the results for the mechanical seals were unexpected. The unbalanced mechanical seal showed lower torque than the balanced O-ring pusher seal. The difference can most likely be explained by the fact that the face profiles are different for the composite seal face of the balanced seal than the monolithic design of the unbalanced seal.
Typical thermal deflections are different for these variations in design. The composite faces tend to have a divergent profile with outside contact, while the monolithic face tends to have a convergent profile with good fluid penetration between the faces. The pressure drop between the seal faces is different, leading to higher effective hydraulic closing forces for the outside contact than for the inside contact. Different spring loads for the designs, which are difficult to set accurately in component seals, would also have a significant impact on contact pressure.
This illustrates two major points. First, specific designs have specific characteristics, and broad classifications are not sufficient to evaluate the power consumption of one type of design. Second, the pressure drop between the sealing interface is critical in determining the actual power consumption of the sealing device. This should be considered with packing and mechanical seals.
The packing friction compares favorably with all of the mechanical seal variants. These unexpected results have led to a reconsideration of the traditional methods for calculating packing friction.
The formula that has long been used to calculate power consumption from compression packing systems is as follows:
P= Pp x RPM x D x µ x Ap x F
P = Power (HP or kilowatts, depending on units used)
PP= sealed pressure
RPM = rotational speed
D = shaft diameter
µ = coefficient of friction between the packing and the shaft
Ap = packing contact area
F = factor, depending on units used
This formula is similar to the one used for mechanical seals, which has been shown to give a good approximation to power consumption levels.