## 2

Unbalance force = 1.77 I !qqQ ) X unbalance (oz • in)

Unbalance = eccentricity of impeller (mm) X impeller wt, mass (g)

Unbalance force = 0.01094 ( HqqQ ) X unbalance (g • mm)

The example also points out that the unbalance force generated by loose fit impellers with keyways mounted in one plane could be quite high. This force could be minimized by staggering keyways or randomly orienting the impellers on the balance mandrel. Shifting of shrink-fitted, well-balanced impellers on multistage and highspeed double-suction pumps after a period of operation can result in unbalance. The shifting of the impeller is due to the relaxation of residual stresses that built up as the impeller cooled and contracted around the shaft. Shaft vibration and flexing tend to relieve the residual stress and cause the impeller to cock or bow the shaft from the original balance centerline.

Standards should be referred to for balancing pumps and their drivers. When balancing, consideration must be given to the need for balancing at rated speed in order to properly evaluate the importance of shaft deflection due to modal components of unbalance. See References 12 and 13.

Hydraulic Unbalance Uneven flow distribution entering the impeller can cause a 1X running speed frequency type of vibration. The unbalance occurs because the flow is not equal in all vane passages. An example of this is a double-suction impeller with a short, straight run to the pump and an elbow in the horizontal plane. Flow from the elbow does not have time to straighten and therefore enters both sides of the impeller unequally. A similar condition results if suction is taken from a tee off the main header. Unequal and unsteady flow into the impeller may cause axial thrust and high axial vibrations. Thus, it is good design practice to install elbows vertically in double-suction pumps.

In double-suction pumps, the nonsymmetrical positioning of the impeller or the offset of the upper case half of the lower case half will cause a 1X unbalance due to nonsym-metrical flow.

Recirculation forces and pulsation recirculation within a pump (which can occur when flow is less than design) may manifest themselves in the form of a noise and/or vibration with random frequencies, along with pressure pulsation that may be seen on a pressure gage. Recirculation may also appear in the piping system as vibration and noise. Increased NPSHA has helped in a few cases, especially if the recirculation is mainly on the suction side of the impeller. After a pump has a recirculation problem in a given system and the flow cannot be increased using a bypass, little can be done to the pump itself unless the system characteristics allow an impeller change.

Antifriction Bearings Vibrations generated by ball bearings cover a wide range of high frequencies that are not necessarily a multiple of the shaft running speed. The frequency readings obtained during analysis are somewhat unsteady because of the resolution of the filters in a hand-tuned analyzer. The amplitude reading may also be somewhat unsteady.

Experience has shown that hand-tuned field analyzers tend to show the last stage of the bearing failure. Monitoring of stress waves or shock pulses (impact energy) on the pump bearing housing will show failure trends that will generally precede an increase in the detectable level of mechanical vibration. This method of failure detection is called acoustic high-frequency monitoring or incipient-failure detection (1FD). A comparison between conventional methods and the acoustic high-frequency method is shown in Figure 23.

Accurate analysis of pump bearings and other machinery can also be made using a velocity sensor good to 1500 Hz or an accelerometer. Data should be recorded and processed through a real-time analyzer with at least 256-line resolution capability and a band selectable analysis option. Analysis of antifriction bearings using a real-time analyzer and equations for calculating frequencies generated by defective bearings can be found in Reference 7.

Accurate and extended analysis of pump bearing vibration is generally not needed. A majority of bearing problems are currently identified by an acoustic noise during operation. What is needed by maintenance personnel is a quick and reliable method to monitor bearings and determine when a bearing is failing and the rate of bearing deterioration. At present, there are some expensive instruments that can be purchased, but none of them meet maintenance requirements.

Currently, the best method for reducing bearings analysis is prevention of the failure. Most antifriction pump bearings fail for one of several reasons: (1) water gets into the oil, (2) automatic oiler is not adjusted properly (this cause is most often overlooked and will continue to produce a short life to failure cycle), (3) product gets into the oil, (4) acidic vapors condense and break down the oil, (5) mounting techniques or fits are improper, (6) new hearing is defective. The solution to high-humidity problems and problems with acidic units is the use of an oil mist lubrication system. If this cannot be economically installed, an aggressive preventative maintenance program on a monthly to every-other-month basis is required.

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