There are hundreds or maybe thousands of equipment in a plant, started from wiring, breaker, ducting, motor, pump etc.If we want to start Condition monitoring the question is what kind of equipment that i can put on my Condition Base Monitoring program.
These criteria below may give you some direction. The condition monitoring are technically feasible if;
Potential failure condition can be detected clearly. This could be increased vibration, temperature, noise, etc.
P-F interval is reasonably consistent. P-F interval is time between Potential failure (point where we know that something is failing) and Functional failure (point where an equipment has failed or unable to fulfill it’s function)
it is practical to have monitoring frequency at interval less than the P-F interval
the net P-F interval is long enough to be some use ( long enough to take action to prevent it from reaching functional failure)
The characteristic vibration of overhung fan is shown in figure 1.
Figure 1. Overhung Fan
Bearing B vibration is related with plane C and bearing A vibration is related with plane D.
Step for onsite balancing for overhung fan are as below:
1. Connect accelerometer in bearing A and perform single plane balancing on plane D or put your accelerometer on bearing B and perform single plane balancing on plane C. See vibration correlation between bearing and plane in figure 1.
2. If the vibration level on bearing A (if you choose bearing A on the first step) are acceptable, move your accelerometer to bearing B and measure vibration.
3. If the vibration on bearing B is not acceptable, then put weight on plane D that equal with trial weight in plane C but 180o apart.
4. Perform static balancing on plane D till vibration is acceptable. Sometime it’s may needed to put equal trial weight on plane C on 180o opposite when you put trial weight on plane D.
Double suction centrifugal pump running at 1000 rpm with oil bath lubrication reported has noisy outboard bearing. Vibration data taken and shows an increase noise floor (figure 1). It was also found a leak on gland packing that cause water spray out toward bearing casing. Increasing noise floor was suspected caused by improper lubrication, this was caused by water get into the bearing. Recommendation was made to drain oil and inspect if the water has enter the bearing.
Figure 1. Vibration spectrum after water contamination
The suspicion that water has entered the bearing was true. They found some amount of water on the drained oil. After flush bearing housing and change the oil, the pump was put back in service and noise floor was decreased. Vibration spectrum after and before oil replacement are shown in figure 2 .
Figure 2. Vibration spectrum after oil replacement
Unbalance is a condition where shaft geometric centerline and mass centerline do not coincide or the center of mass does not lie on the axis of rotation.
Vibration spectrum will be dominated at 1x running speed and the time waveform will sinusoidal for pure unbalance.
There are three types of unbalance condition :
1. Static unbalance : In this type of unbalance we have a heavy spot at a single point in the rotor, it will shows up even when the rotor is not running or if you put it on the frictionless bearings the rotor will turn so the heavy spot is at the lowest position.
The vibration signal at each end of the machine in the same direction will be in phase. There will be 90 o+ 30 o phase different between horizontal and vertical direction.
Figure 1. Static unbalance
2. Couple Unbalance : A rotor with couple imbalance seems well balanced in static condition (the rotor not turn when placed on frictionless bearings). But it going to produce centrifugal force on the bearings when the rotor is rotating.
3. Dynamic Unbalance : is a combination of static and couple unbalance. Spectrum at 1x running speed will dominated overall vibration. The highest vibration usually at horizontal direction where the machine can move more freely. Vibration signal at each end of the machine taken at the same direction will be 0 o to 180 o. There will be 90 o+ 40 o phase different between horizontal and vertical direction.
Misalignment is a condition where the centerlines of coupled shafts do not coincide. If the misaligned shaft centerlines are parallel but not coincident, then the misalignment is said to be parallel misalignment. If the misaligned shafts meet at a point but are not parallel, then the misalignment is called angular misalignment. Almost all misalignment conditions of machines seen in practice are a combination of these two basic types.
Type of misalignment :
1. Parallel Misalignment : condition where the misaligned shaft centerlines are parallel but not coincident. This type of misalignment produce shear forces and bending moment on the coupled end of each shaft. Vibration are dominated at 1x and 2x running speed at radial direction and most often spectrum at 2x lf are higher. Axial vibration at 1x and 2x running speed will be low.
Figure 1. Parallel misalignment
2. Angular Misalignment : condition where the misaligned shaft meet at a point but not parallel to each other. This misalignment produces a bending moment on each shaft. And generate high vibration at 1x and some vibration at 2x and 3x in the axial direction at both bearings. The vibration will be 180 degree out of phase across the coupling in the axial direction, and in phase in the radial direction.
Soft foot is a condition where you have uneven level of feet. This cause distortion on the machine frame as you tight or loosen the bolt. Effect of soft soft foot could be damaging for an equipment or at least decrease it’s operating life. Softfoot or casing distortion will cause several problems like :
1. Internal misalignment :Casing distortion will cause misalignment between the internal bearings. As the machine frame distorted due to softfoot, the bearing move relative to each other and when this relative movement exceeds bearing tolerance it going to deflect the shaft . Every time the shaft turn there will be back-and-forth bending moment, after a period of time, the shaft may develop fatigue crack and fail.
2. Distorted Bearings : As the machine distorted, the shaft bearings are also distorted. The effect of this distortion, the bearing will have two load zones 180 o apart. Softfoot is one of the major causes of outer race distortion in properly installed bearing.
Softfoot problem on a motor can be identified by the appearance of spectrum on 2x lf. This is caused by uneven air gap between rotor and stator due to frame distortion.
1. Units of measurement : Rotary lobe blower vibrations are measured in inches/sec. Measurements of spike energy is not recommended for judging blower condition because the rotary lobe blower has inherent impacting bearing loads.
2. Measurement location : vibration should be measured at bearing locations on the housing
These are guideline for assessing rotary lobe blowers rigidly mounted on the stiff foundations.
If the blower is operating at "review required" levels then the installation must be fully evaluated to determine the source or cause of vibration and the cause shall be corrected.
In general, blower vibration levels should be monitored on a regular basis and the vibration trend observed for progressive or sudden change in level.
Element Generating Vibration in rotary lobe blower :
1. Blower Inherent characteristic a. Impacting bearing loads excite component/system natural frequency. b. Pressure pulsations set up vibration at four times running speed
2. Rotary lobe blower has very close clearances between the impeller and the housing. The impeller contact will setup vibrations as follows : a. Impeller to impeller frontal lobe contact-if contact is between only one set of lobes, the vibration frequency will be 1x rpm, if both sets of lobes contact, the vibration frequency will be 2x rpm. b. Impeller to cylinder contact-the vibration frequency will depend on the number of impeller tips contacting the cylinder which could range from one to four times the RPM. c. Impeller to head plate contact-the vibration frequency will be erratic and unsteady.
3. Damaged gears will generate vibrations at gear mesh frequency (GMF)
4. Damaged gear will generate vibrations at bearing frequency (BPFI,BPFO, FTF, BSF)
5. Rotor unbalance and bent shaft will generate vibrations at 1x rpm.
6. Blower/driver coupling misalignment will generate vibrations at 1x RPM and 2xRPM
7. Acoustic resonance in the blower inlet/discharge piping will generate vibrations at 4x RPM.
8. Operation of rotary blower at or near system torsional may cause impeller lobe contact and increase vibration
9. External piping if not properly isolated will transmit vibrations into the blower
10. Foundation design and method of mounting has considerable effect on blower vibrations.
This was happened on 15 kW motor that drive three stages centrifugal compressor. During routine measurement there where no evidence of any problem on the motor (figure 1).
figure 1
22 July 2010, compressor was tripped because motor was overloaded, mechanical inspection found leaks at compressor cooler.During the inspection period, motor solo run test was done to make sure that there is nothing wrong with the motor.
Motor was noisy, like something was rubbing especially at NDE position. TWF show different pattern than routine measurement (figure 2).
Figure 2.
It was decided to open bearing cover and re-grease the bearing.We found that bearing housing cavity was filled with grease. We recommend to clean bearing cavity and re-grease the bearing. The vibration spectrum after re-greasing is shown on figure 3 below :
Shop balancing was conducted after impeller repair. Running test was carried on the result shown on table 1 was amazed everyone.
The higest vibration was 32 mm/s rms. Dominated frequency was 1x running speed, the question was “ are they really balanced this equipment?”.
Balancing document review shows that the shop consider the impeller as centerhung fan, although residual unbalance was still on tolerance for grade 2.5.
Then it was decided to rebalanced the impeller (on site), and the final result vibration drop from 33 mm/s to 2.3 mm/s.
Note : There are different treatment in balancing procedure for centerhung and overhung rotor. Failure on recognition will cause unacceptable balancing result.
There some rules of how to mount the transducer to ensure you get representative data : 1. Vibration readings must be taken on a surface stiff enough that it is not affected by the pressure exerted by the probe or accelerometer. 2. All vibration reading should be taken with the transducer perpendicular to the surface of interest. 3. Vibration signals containing high frequencies must be taken with the transducer tightly screwed to the surface, as hand pressure can not hold it tightly enough to the surface for it to correctly follow high frequency 4. Magnetically mounting the transducer is preferable to a hand-held reading, but not as good as a hard mount as far as high frequency response. The magnet will rock if put on a curved surface, reducing its usable frequency range.
Operation department reported increasing condenser vacuum. Condenser vacuum should be maintained at -710 mmHg. In the last three days, the highest vacuum was only -691 mm Hg. This condition was causing decreased turbine and overall power plant efficiency.
Condenser vacuum on this power plant were maintained by one vacuum pump ( one vacuum pump as redundant). The analysis was focused on this equipment. Infrared camera was used to identified any abnormality on the piping that connect condenser with this equipment. And it was found that some portion of the suction piping was 8 oC cooler than the other section.
The coolest section was after suction valve. It was suspected that the suction valve was blocked or partially opened and causing restricted flow of gasses.Recommendation was made to inspect suction valve opening/condition.
Inspection by mechanic verify that suction valve was only partially opened. Infrared scanning after inspection show the temperature was distributed evenly on all section and condenser vacuum rise to 705 mm Hg.
Predictive maintenance is based on comprehensive measurements of mechanical condition. The ability of minimizing breakdown failure and giving direction to maintenance action and repair are the benefits of Predictivemaintenance. From the financial perspective Predictive maintenance is less costly compared with preventive maintenance and corrective/reactive maintenance. A major return in predictive maintenance result in maximizing production availability and minimizing the risks, cost and amount of production lost resulting from unexpected shutdown or failure. Preventive maintenance also known as time base maintenance is type of maintenance that scheduled at specified time interval, it doesn’t matter whether the equipments are showing sign of defect or not . This kind of maintenance required inspection and sometimes replacement of some parts. Thus , preventivemaintenance typically result in unnecessarily replacing components in good condition. The worst part is when machine that in good condition disassembled for preventive maintenance and returned to service in poorer condition. Reactive maintenance usually result in “fire fighting” where the problem causing failure are never corrected . Additionally outright may also lead to secondary failure. For example, misalignment problem could lead to bearing damaged which, in turn, causes costly damage to the bearing housing and shaft and possibly unscheduled shutdown.
NEW BEARING FAIL 2 DAYS AFTER INSTALLATION CAUSED BY IMPROPER BEARING FIT.
During power plant annual inspection, PdM department gave a recommendation to replace motor bearing and repair bearing housing dimension. The job was done by third party in a week and a acceptance test (no load condition) was conducted . Vibration spectrum and PV from acceptance test is shown in fig 1.
Fig 1. Vibration spectrum and PV from inboard bearing from acceptance test
The data was taken at 7 July. From vibration spectrum, it was noted some activity at high frequency which usually corresponded to bearing defect, but since it’s magnitude was still lower than acceptance limit criteria it was decided to pass and put it into operation.
9 July , operator reported very high noise from this machine. Vibration spectrum and PV is shown in picture 2
Fig 2. Vibration spectrum and PV after two days in service
It was noted increased noise floor at vibration spectrum and PeakVue. BPFO and BPFI also appear on vibration spectrum. Then it was decided to draw it from service and install new bearing. During inboard bearing disassembly, it was found that bearing housing diameter was 0.5 mm smaller than it should be. Heavily marked path pattern was identified in raceways. The probable cause was smaller dimension of bearing housing make bearing internal clearance get to tight and cause high load at he bearing. This is way the bearing only last for 2 days.
Mar 31, 2010
Dynamic Absorber Aplication on Condensate Pump (Vertical Pump)
Power generation with rated output 400 MW operate two condensate pumpto transfer water form condenser hotwell to low pressure heater. Theese pump are 6 stages vertical pump. If one of the pump is out service it would derated 50 % from the design output. The plant operator reported high vibration one of the pump. We were invited to conduct vibration analysis. Figure 1 shows machine diagram and picture :
Figure 1. Machine diagram
Vibration data ( table 1) were taken at motor outboard an inboard bearing.
Vibration
Motor
Pump
mm/s rms
Out
In
In
Out
Horizontal
7.0
2.5
0.8
--
Vertical
1.9
1.3
0.6
--
Axial
0.6
--
0.3
Table 1. Vibration data
From table 1 it was noted that vibration were dominant at horizontal direction ( peroendicular with pump discharge ) and the dominant frequency was at 1x order as shown in figure 2. Amplitude ratio between horizontal and vertical nearly 3 : 1.
Figure 2. Vibration spectrum at Outboard bearing-horizontal
Our hypothesis was we have resonance problem at horizontal direction, bump test also indicated that the pump natural frequency was very close with pump operating speed. To avoid resonance condition we have to move pumps natural frequency but this would cause the pump out service and decreased plant output. The only available option was to install Dynamic Vibration Absorber. This was made form rectangular bar and a mass adjusted to get natural frequency equal with pump natural frequency. Table 3 show vibration amplitude after absorber installation
Vibration
Motor
Pump
mm/s rms
Out
In
In
Out
Horizontal
2.1
0.9
0.6
--
Vertical
1.9
0.9
0.6
--
Axial
0.8
--
0.3
Table 3. Vibration after Absorber istalation
Figure 3. Vibration spectrum after absorber installation
After absorber installation vibration dropped from 7 mm/sto 2.1 mm/s.
