ROTATING EQUIPMENT SCIENCE

Tuesday, November 17, 2020

A CASE STUDY OF ABNORMAL SOUND FROM VFD MOTOR

Variable speed drives have always raised challenges for the condition monitoring personnel due to its varying speed and load conditions and the necessity of fixing the operating parameters before data collection or use the techniques such as order tracking etc. to diagnose the faults. Here we are discussing an interesting case, where an abnormal sound, similar to a damaged bearing or rub sound, arose from the motor during the commissioning run, after being a standby for more than 6 months. Different tests were performed before concluding the results, which are explained below. The motor was in continuous service until stopped a few months back due to the decline in demand.

The subject motor drives a single stage radial flow blower.

Observations

Abnormal sound noticed from the motor when operated in the minimum speed range during low load. Vibration spectrum collected from motor showed raised noise floor, around and above 270 Hz up to 3500 Hz, the majority of the peaks observed above 2000 Hz range.

Pic.1 Vibration spectrum

Run-up test was done to identify the changes in peaks. The speed was increased from 3150 rpm to 3490 rpm. It can be seen that some of the noise below 1000 Hz being reduced with increase in speed.

Pic.2 Run-up test cascade plot

The coast-down test was performed to rule out the electrical noise from the vibrations.

Pic.3 Coast-down cascade plot

During the coast-down or ramp down, it was observed that the peaks above 2000 Hz immediately vanished. A detailed study was done to identify the cause of these peaks. For understanding that the working of VFD is briefly described below.

Working of a VFD

VFD module consists of input diode rectifiers, DC bus and IGBT inverter. The AC current supplied to the VFD module is converted into DC current by diodes, which is smoothened out by the capacitor of DC bus. The smoothened DC current is converted into the AC current of required frequency by pulse width modulation, achieved by using the IGBT.

Here, I am not going to discuss in detail about the working of input and DC bus sections. Our concern is regarding the output current and what component of this current could be leading to the vibrations in the motor. To understand this, let us discuss how IGBT works.



Pic.4 VFD module

As mentioned, IGBT or Insulated Gate Bipolar Transistors varies the frequency by pulse width modulation. The pulses are developed by simply “opening” and “closing” the circuit, just like by using switches. Similar to normal switch operation, when the switch is closed, the current flows, when the switch is open, the current flow stops. Now, to control the width of the pulse, the timing of each pulse have to be varied. When the pulse width is more, the voltage output is more and vice versa. Hence, to obtain a sine wave output, the pulse width at the centre of each half-cycle will be more than that at the ends. The timing of each pulse is varied by controlling the opening and closing of transistors.

If you remember the basics of the transistor, which consists of collector, emitter and gate terminals, where the main current passes through the collector and emitter (which is built as normally open or normally closed) and the gate signal triggers the connection between the two (collector and emitter), it will be much easier to understand (If you do not have the basic idea, I strongly recommend to refer some articles on transistor working). The chip, which provides the control timings based on the program, activates or deactivates the current flow through the transistor. Thus, we get output pulses.

Pic.5 Pulse width modulation and generation of the voltage wave

We have an idea now on how the voltage is varied (i.e. by varying the pulse width). However, the frequency is varied by increasing or decreasing the time interval between each pulse. As the time between the pulses is increased, the frequency decreases as time decreases, frequency increases. (Marking this as important as it explains how our abnormal sound is appearing)

Now, getting back to the output current, as discussed before, the output sine wave is not a pure AC sine wave but contains many pulses, which varies in width. The harmonic distortion in the wave depends on the number of pulses per cycle and the gap between each pulse, as described in the pic. 6 below. Note that the same 20 pulses per cycle are used here and only the timing is varied in comparison to pic. 5. (Excuse me for the rough hand sketches)

As the number of pulses is increased, the harmonic distortion decreases and vice versa.

Pic.6 Varying frequency by controlling the timing between each pulse

The number of pulses per second of the IGBT output (the carrier wave) can be any of the following, 2 kHz, 4 kHz, 8 kHz or up to 16 kHz. The more number of pulses, the better the output. However, as the number of pulses are increased, the heat developed in VFD will be high and requires sophisticated cooling and thus the requirement of a larger size VFD module.


Coming back to the motor sound, in our case, the switching frequency was identified to be 2 kHz, which is observed in the spectrum. The raised noise floor was an indication of excited natural frequencies, probably the stator lamination or winding resonance at the carrier frequency. Hence, recommendations were provided to check and increase the switching frequency to 4 kHz (Can be done only in limited models) or to change the speed control from scalar mode to Direct Torque Control (DTC) mode (which was available in our ABB ACS880 model). DTC mode has several advantages over the conventional V-HZ (scalar) control, due to the absence of separate voltage and frequency controlled PWM. Approximate sinusoidal flux and current can be achieved through DTC. The recommendations were implemented and the audible sound vanished. Note that we are just changing the exciting frequency and shifting it away from the natural frequency of laminations. In our case, we reduced the distortions and thus the resonance amplitude. In case we increase the frequency, the generated tone is shifting above our audible range and may still be observed in the frequency spectrum. Even before, during the test, when the speed was increased (increasing the frequency of current), the audible sound slightly decreased, thus confirming the doubt.


So, when you hear an abnormal sound from VFD motor next time, do not rush into conclusions. Take your time, do the tests and confirm the actual problem. The same switching frequency was supplied to similar motors in the plant, but only this particular motor produced the abnormal sound. It was noticed that only this motor utilized scalar control mode where other motors operated in DTC mode. Hence, our second recommendation was implemented, which reduced the sound.




Wednesday, November 4, 2020

HIGH VIBRATION OF A VERTICAL TURBINE PUMP MOTOR DUE TO STRUCTURAL RESONANCE

INTRODUCTION

High 1x vibration observed at motor NDE. The vibration was predominant in horizontal direction compared to that in vertical. Structural resonance was suspected. Cascaded plot data were collected during coast-down and bump test done on the structural components and motor body to confirm the NF. The level of the motor stand also checked and was found tilted by 6 mm/meter in the horizontal direction.

HISTORY

The pump had a history of high vibration and was mostly kept as standby. The usual vibrations were around 8-11 mm/s at motor NDE. During service, the flow rate of the pump dropped suddenly due to foreign object entrapment inside the casing. The pump had to be pulled out for the overhaul. At the same time, the motor NDE bearing which was having abnormality was replaced. But, after the overhaul activities, the vibrations found very high, which is explained below.

High directional vibration, predominant in horizontal (less stiff) direction was observed during the trial run. Even during the decoupled solo run, the motor vibrated above 11 mm/s. Resonance was suspected and coast-down test was done during the solo run to confirm the existence of resonance. The plots are provided below for reference.


Radial horizontal (less stiffness direction) vibrations during load trial (21 mm/s)

Radial vertical (greater stiffness direction) vibrations during load trial (1.3 mm/s)

Coast-down test data during solo run

Trend plot of 1x during coast-down

From the plots, it is clear that within a 20 rpm speed change, the vibrations dropped from 9 mm/s to 2.6 mm/s, during the solo run.

The load-trial vibrations were 21 mm/s, which could have damaged the motor bearings.

Bump test was done to identify the resonating structural component. The motor stand natural frequency in horizontal and axial mounting faces found coinciding with running speed frequency.


Bump test plot - Motor stand horizontal indicating peaks at running speed


Bump test plot - Motor stand mounting face (axial) also indicating peak close to running speed

Based on the above test results, it was concluded that the motor stand was weak in the horizontal direction and the mounting faces required additional support or rigidity. Hence, braces and gussets were added to improve the rigidity of mounting flanges and in the horizontal direction. Also, the motor stand level was corrected to within 0.5 mm/meter as per the equipment manual.


Suggested locations for braces and gussets on motor stand

Suggested locations for braces and gussets on motor stand


After improving the rigidity and correcting the level, the vibrations came down to 2 mm/s RMS in the horizontal direction.


Vibrations spectrum after implementing the corrective actions


SUMMARY

Improvement: Vibrations reduced from 21 mm/s to 2 mm/s at motor NDE horizontal.

Impact: Will improve motor and pump bearing life, thus saving the maintenance cost.

Activities performed:

1.       Improved structural rigidity by providing braces and gussets.

2.       Maintained the motor stand level as per the OEM standards.

Abbreviations used:

NDE: Non-Drive End/ Outboard

NF: Natural Frequency

RMS: Root Mean Square (value of vibration)

OEM: Original Equipment Manufacturer


Sunday, November 24, 2019

BELT DRIVES: INSTALLATION, TENSIONING & ALIGNMENT

Drive belt failures are one of the most common problems in industries. Premature failure of the belts is caused by improper installation, incorrect pulley alignment and improper tensioning of the belts. Hence, it is discussed here on how to properly install, align and tension the belts.

Belt Installation:
For easy installation, some people use screwdriver or pry-bar to push the belt(s) (known as rolling or prying) into the pulley grooves. This is done to not disturb the alignment. But, in doing so, the belts will get damaged. The highest quality belts are said to lose 50% of its initial tension during operation. So, if we are using improper methods like screwdriver/crowbar in pushing the belt, it will further reduce the tension, due to the plasticity of rubber. Also, it might damage the fiber cords inside the belt which leads to a lower belt life.
The proper method will be to bring the movable equipment closer to the fixed one, to make the installation of belts easier with hand. After placing the belts properly in the pulley grooves, bring back the movable equipment to the original position and adjust the belt tension.

Belt Tensioning:
To reduce the maintenance time, some people use "feel method" for belt tensioning. As per the study conducted by Bestorq, only 1% of such belts are tensioned correctly. Proper belt tension is an important factor in determining the belt life as well as the bearing life. High belt tension will induce greater stress on bearings and shaft, whereas low belt tension leads to excessive slip and temperature rise, both of which affect the bearing and belt life.
The belts can be tensioned properly using a pencil type tension gauge or by using a belt frequency meter. Pencil type tension gauge is the cheapest device for tensioning and can be purchased for approx. USD 40. The tension gauge uses force per deflection method to identify the belt tension. The recommended force for a belt type will be provided by the manufacturer or will be available on their website. The deflection for standard V-belt is generally calculated as 1/64 inches per inch of span length.
Allowed deflection for V-belts = (1/64 ") * Span length in inches


The calculated allowed deflection value is set in the pencil type tension gauge by moving the rubber washer. The small washer on the plunger indicates the force value. Initially, it is kept at zero. The gauge should be kept at the center of the span length, on the tight side of the belt and force is applied until the deflection value set in the gauge comes in-line with the other belt surface or in case of a single belt, the initial belt position marked by using a straight edge or other methods. This force value should be within the limits provided by the manufacturer.


By using a belt frequency meter, the values such as belt span and mass data (which can be availed from manufacturer) should be entered and the sensor is kept as close to the centre of the span. The longer edge of the sensor is kept parallel to the belt axis (along the belt span) and with a 1/2 inch gap from the belt. Tap the belt with some tool to make it vibrate. The meter will show the belt tension in units of force. In short spans, these instrument tends to give wrong readings. So, it is always good to check with a calibrated pencil-type tension gauge.


Pulley Alignment:
Pulley alignment is one of the major factors determining the belt life. Incorrect alignment leads to uneven tension across the belts and will lead to the flapping of belts with the lowest tension. The belts which flaps might turn-over and lead to failure. While the other belts have excess tension and these fail quicker than normal.
Generally, pulley alignment is done using straight edges for a smaller centre to centre distance or by using thread for a longer span. But the precise method will be to use laser alignment tools such as SKF belt alignment tool TKBA 40, Easy-Laser etc.
Always remember, the grooves are supposed to be aligned and not the sheave surfaces. Hence, it is always better to measure the rim thickness. If the thickness is different in both pulleys, then care must be taken to make the grooves aligned. Because of the same reason, it is called as belt alignment.



BANDED BELTS: PROS AND CONS

Banded belts are multiple belts joined together by a band to form a single drive unit. This is useful where there are pulsating loads and the belt flaps/vibrations will be more. This tends to twist the belts during operation. Pulleys with a longer centre to centre distances are prone to excessive belt flaps and banded belts are very useful in such cases.
In most industries, excessive belt flaps are observed in drives with multiple belts, due to pulley misalignment. Mostly, few belts out of these, which are under more tension, will fail first. Thes failed belts will get turned over and damages the other belts and sometimes leads to equipment or bearing failure. These problems can be avoided using a banded belt.



Listed below are the advantages of banded belts:
  1. Prevents damages and jumping off the sheaves due to excessive belt flaps.
  2. Uniform load distribution across the belt.
Some of the disadvantages include
  1. Dirt and debris cannot escape the bands. Hence it is not suitable to be used in dirty environments.
  2. Pulley wear must be kept to a minimum, or else the top band might rub with the rim top surface and gets cut.
Hence, if you have a clean environment and is facing issues with frequent belt failure (especially turning over of belts), try the banded belts. Keep in mind that the pulley groove wear must be minimum. Try it with a new set of pulleys, if possible.

Monday, August 5, 2019

MINIMUM SUBMERGENCE REQUIREMENT IN PUMPS

Minimum submergence, as the term suggests, is the minimum depth of pump inlet bell or minimum distance to water surface required from the inlet bell of a pump. This is very important to prevent the formation of vortices during operation.
If the submergence is reduced, initially water swirl is formed. Gradually, as the level comes down, surface vortices are formed. Again as the level decreases, strong vortex with a dye core is formed, which induces turbulence at the suction. Lastly, vortex with a hollow/air core is formed sucking air inside. This might lead to the loss of efficiency of the pump and cavitation due to the presence of air.
Pumps are made to handle liquids only. Any air entrapment will reduce pumping efficiency. If the volume of air is large, the pump stops pumping. To remove air from the pump casing and suction line, generally, we do priming before starting a pump.

The term is of greater importance while designing a pump intake as per Hydraulic Institute standards. In case of designing a river or large water body intake, the minimum water level recorded at the site during the driest seasons (sometimes, forecasting of annual decrease in water level will help) to be considered to calculate minimum submergence.

Minimum submergence S is given as,

S/D = 1 + 2.3 Fd

S = Minimum Submergence
D = Inlet Bell Diameter
Fd = Froude's number = V/sqrt(gD)
V = Velocity of flow at suction (can be determined from flow rate, Q of the pump and inlet bell diameter, Q=AV)
g = Acceleration due to gravity = 9.81 m/s2

Effects on Pump
  1. Vortex formation induces turbulence flow through the pump, impacting the pump bearings.
  2. Air entrapment leads to cavitation, reduction in pumping efficiency and sometimes loss of prime.

Saturday, July 6, 2019

INITIAL STAGES OF BEARING FAILURE PRODUCES HIGH FREQUENCY VIBRATIONS? HOW?

To understand this, let's take an example of a simple pendulum.
We know that time period of oscillation of a simple pendulum is given as,
T=2π√(L/g)
Where T is the time period of oscillation
L is the length of the pendulum
g is the acceleration due to gravity
From the equation, it is clear that the time period is proportional to the length of the pendulum.

From the figure, it can be seen that for the same angle or force, depending on the length, the amplitude or oscillation length changes. The amplitude is directly proportional to the length of the pendulum which in turn is proportional to the time period of oscillation.
i.e.; for a shorter amplitude or length, the time period will be less and for larger amplitude and length, the time period will be more.
The frequency is reciprocal of the time period, 1/T.
Hence, for small amplitudes, the time period will be less and frequency will be more.

Now, coming back to the bearings, initial failures are microscopic, the damage gap of which will be in nanometers or microns (just like a shorter length pendulum) which produces small amplitude vibrations. This, in turn, generates very high frequencies. As the failure progress, the damage becomes macroscopic, with the damage gap visible to the naked eye, producing high amplitude vibrations at lower frequencies.

In a more detailed way, as the damages on bearing are microscopic, the forcing component (ball, or races) will pass through the defect quickly, making the contact time small and frequencies high. As the damage grows, there will be more contact time and hence frequencies becomes lower. When the damage is severe, the contact time will be more and impacts are produced, leading to raised noise floor in vibration spectrum.

HOW POLE PASS FREQUENCY IS SLIP SPEED TIMES THE NUMBER OF POLES?

In vibration analysis, pole pass frequency is one of the important terms used during analysis of induction motors.
Pole Pass Frequency = Slip Frequency X No. of Poles of Induction Motor

Generally, many have doubts on why Pole pass frequency is Slip frequency times the number of poles in an induction motor and not running frequency times the number of poles. The answer can be better understood through the induction motor working principle.

How induction motor works?

In an AC induction motor, which is generally used in industries, the alternating current is supplied only to the stator windings. This produces an alternating flux field which rotates at the synchronous speed (speed of magnetic flux). The flux produced by the stator induces a current in the rotor windings as per Faraday's law of electromagnetic induction. The induced current in the rotor will also produce a flux around the rotor coils which lags the stator flux. The induced current in rotor tries to oppose the cause of its production as per Lenz's law, which is the relative speed between the stator field and rotor. This makes the rotor or armature to rotate and to try and catch up with the Rotating Magnetic Field (RMF) of the stator. But once the armature speed matches the RMF, there is no relative speed between RMF and armature, thus decelerating the armature due to mechanical losses, which in turn develops the relative speed again and the same process continues. Hence, always the armature speed will be less than RMF or synchronous speed, giving rise to slip.

Now back to our topic,
If we consider rotor or armature as stationary, the RMF can be seen to be rotating at slip speed. Hence the poles will be passing armature at a slip speed and not the running speed.
Therefore, pole pass frequency is the number of poles times the slip speed.

Use of pole pass frequency

1. To detect eccentric rotor with varying air gap: Increased amplitude at 2x Line Frequency with pole pass sidebands can be seen along with 1x vibrations.

2. To detect cracked or broken rotor bars: Harmonics of 1x with pole pass sidebands are observed in spectrum along with beating. 

Monday, May 27, 2019

MECHANICAL SEALS FOR CENTRIFUGAL & ROTARY PUMPS (AS PER API 682)


We know that rotary/centrifugal pumps are used to develop head and discharge liquid at a specified flow rate. The medium to be pumped will be enclosed in a casing. The rotor (impeller, vanes, screw, lobes etc.) will impart the required dynamic motion to the liquid.
The liquid inside casing should be contained in order to prevent leakages, which might reduce the efficiency, cause environmental problems etc. The gaps between two static parts can be sealed by using gaskets of appropriate size. But the gap between shaft and casing cannot be sealed in this way. For this purpose, we use mechanical seals.
Here we are discussing the types of mechanical seals as per API 682 and some common terms related to seals. API 682 is a standard developed for the mechanical seals used in centrifugal/rotary pumps.

TERMS

  1. Barrier Fluid: It is a pressurized fluid used in dual mechanical seals or double seals to isolate the process fluid from the environment. It also acts as a cooling and lubrication medium for seal faces. The pressure of barrier fluid is higher than that of the process fluid.
  2. Buffer Fluid: It is a fluid used in the dual mechanical seal which is supplied at a pressure lower than the process fluid pressure. The main purpose is for lubrication.
  3. Flush: This is the fluid introduced at process fluid side for cooling and lubrication of seal faces. This is found in single seals. It is generally provided very near to the seal faces.
  4. Quench: It is a neutral fluid introduced at the atmospheric side of the seal, used to prevent the solid formation that might damage the seal faces. Generally, water or steam is used for this process.
  5. Throat bushing: It is a restrictive close clearance provided around the sleeve/shaft between the seal and impeller. A Teflon or graphite bush provided on the back plate serves this purpose.
  6. Throttle bushing: It is a restrictive close clearance provided around the sleeve/shaft at the outboard/atmospheric end of mechanical seal gland. A Teflon or graphite bush provided on the back plate serves this purpose.
  7. Gland: It is the seal housing with threaded holes to provide a quench/flush connection.
  8. Stationary face: The face of the seal which does not rotate along with the shaft. These are generally mounted on the gland and/or backplate, depending on the type of seals. Cartridge type seals come with an integral assembly of both stationary and rotary faces.
  9. Rotary face: The face of a seal which rotates along with the shaft. These are mounted on shafts.
  10. Pusher type seal: The secondary the seal is mechanically pushed along the shaft or sleeve to compensate for the face wear.
  11. Non-pusher type seal: The secondary seal is fixed to the shaft.
  12. Internal circulating device/ Pumping ring: It is located in seal chamber, mounted on the shaft, used to circulate barrier fluid through cooler or buffer fluid reservoir.


Required Minimum Flush Flow rate and Pressure in Seal

Minimum required seal chamber pressure for inner unpressurized dual seals and single seals is 3.5 bar (50 psi) or 10% above maximum fluid vapour pressure at seal chamber fluid temperature.
Minimum seal flush flow rate is 8 L/min (2 gpm) per seal.

API 682 coverage based on temperature and pressure range

Temperature range: -400C to 2600C (-400F to 5000F)
Pressure range: 0 bar to 34.5 bar (0 psia to 515 psia)

PARTS OF MECHANICAL SEAL

  • Seal faces: As explained above there will be rotary and stationary faces. Ideally, the faces will not come in to contact with each other, as the process fluid or barrier/buffer fluid acts as a lubricant to separate them. The material of construction will be Teflon, graphite/carbon, Silicon Carbide (SiC), Tungsten Carbide(TiC). These are used in different combinations based on the service.
  • The seal face mating surfaces are lapped to a high degree of precision to maintain the flatness.
  • Retainers: These are metallic parts used to hold the rotary seal faces. These are provided with ant-rotation slots/pins to prevent the seal faces from rotating.
  • Fluoroelastomer/ perfluoroelastomer: These are the O-ring materials used for different service temperature. These are provided to prevent leakage through shaft/casing and seal face joints. It is like a gasket for the seal faces. Fluoroelastomers are used at normal operating temperature. Viton material belongs to this category. Perfluoro elastomers are used in high temperature and chemical services. Kalrez and Chemraz materials belong under this category.
  • Springs: These are found in pusher type seals used to mechanically hold the seal faces in position and maintain the uniform gap between seal faces irrespective of wear and minor pressure fluctuations.


SEAL PLANS

Single Seals
Plan 01
Plan 02
Plan 03
Plan 11
Plan 12
Plan 13
Plan 14
Plan 21
Plan 22
Plan 23
Plan 31
Plan 32
Plan 41
Plan 51

Dual/Double Seals
Plan 52
Plan 53A
Plan 53B
Plan 53C
Plan 54
Plan 55

Quench Seals
Plan 61
Plan 62

Single Seals with Leakage collection/detection
Plan 65A
Plan 65B
Plan 66A
Plan 66B

Secondary Containment Seals
Plan 71
Plan 72
Plan 75
Plan 76

Dual Gas Seals
Plan 74

Engineered Piping Plan not defined by any of the above mentioned
Plan 99

Thursday, May 9, 2019

MOTOR CASE STUDY-1: INCREASED CLEARANCE IN BEARING AND ROTOR RUB


A rubbing sound was observed from the compressor cooling fan motor. Vibration readings were collected only from the motor as the fan is directly mounted on the motor shaft. The obtained spectrum and waveform are given below.


An increased acceleration value in time signal indicates heavy impacts and asymmetrically shaped waveform (greater amplitude at positive side compared to the negative side) indicates restriction in rotor movement possibly created by rubbing. And also rubbing could happen only when there is increased clearance in bearings.
The highest peak is observed at approx. 5.4X orders and its harmonics also found in the spectrum. Since it is a non-synchronous peak, the possibility is bearing damage. Sidebands (sideband frequency is 18Hz, possibly from a damaged ball passing cage) also present with raised noise floor, confirming bearing damage.

Recommendations were provided to replace the bearings and inspect motor armature. The bearings were found having play and damage. Minor rubbing observed on the armature, following which winding resistance and other electrical healthiness checks were performed, ensuring the satisfactory condition of the motor. The bearings were replaced and the motor was reused.

Tuesday, April 2, 2019

PUMP CASE STUDY-2 HIGH VIBRATIONS IN HORIZONTAL PUMP DUE TO PIPE STRAIN

An OH1* pump of Superflow make failed frequently. High 1x RPM vibrations were evident in the spectrum revealing unbalance/base looseness issues. The pump was added into the bad-actor list due to the failures.
18mm/s vibration at 1x RPM
Pump details:
Speed: 3546 RPM
Type: *OH1 = Overhung, foot mounted pump as per API 610
Service: Ejector water circulation
Flow: 152 m3/hr
Head: 51 mlc
Seal: Single seal plan 11,62
Impeller type: Closed

Observed damages:
  1. Mechanical seal damaged.
  2. Bearing failure.
  3. Broken shaft.
  4. Worn out impeller nose and casing wear ring.

A detailed inspection revealed cracks on the base frame below the pump led to the high vibrations. Furthermore, pipe strains on the pump induced by improper pipe supporting led to the high vibrations, which in turn led to the cracking of the base frame.
The shaft failed at an area of reduced cross-section. The inspection revealed that the corners were sharp and not rounded leading to a high-stress concentration at that area, eventually leading to failure.
Cracked frame
Cracked frame
Corrective actions:
  • On Pump
    • Seal replacement.
    • Bearing replacement.
    • Trimming of impeller nose.
    • Fabrication of new casing wear ring.
    • Fabrication of new shaft and rounding off (fillet/chamfer can be provided) sharp corners.
  • At site
    • Correcting pipe strain by providing proper pipe supports.
    • Weld repair on the cracked base frame.
    • Alignment.
After correction, vibrations are down to 3.569mm/s at 1x
After correcting, the vibrations came down from 18mm/s to 3.569mm/s at 1x.

How it was solved?
  • Eliminated pipe strain at suction and discharge nozzles by providing proper pipe supporting.
  • Repaired the base frame cracks by welding.
  • Ensured proper coupling alignment.