ROTATING EQUIPMENT SCIENCE

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.

Saturday, March 30, 2019

DRYER ROLLER CASE STUDY 1 - HIGH IMPACTING SOUND FROM DRYER SUPPORT ROLLERS


One of the most challenging cases came when one of the product dryers started creating high impact sounds from the non-drive side/outlet side support rollers. The equipment is a big rotary drum supported by 2 rollers each at inlet and outlet side. The drums are fitted with a replaceable component called rider rings one each at inlet and outlet side, which rests on the two support rollers. A slope of 1 deg. is maintained so that the dried product moves down by gravity. Hot air is blown through the drum, to facilitate the product drying. Gland packings are provided at inlet and outlet sides to prevent hot air leakage. The drum is rotated by a motor through gearbox, which transfers power through a spur gear arrangement provided on the drum. Many suggestions came from the maintenance like it is because of eccentricity in drum rotation, bearing damage etc. Due to high market demand, stoppage of equipment without identifying the root cause would lead to severe wastage of time and will incur huge losses. But the correct problem was identified by thorough observation and teamwork.

IMPACTS
Impacts are large forces acting over a short time, which can lead to fracture failure/breakage of the impacted component.
In this case, large hammering sounds were observed from the outlet support rollers. It was very difficult to identify the cause of the problem due to access restrictions. The whole support roller assembly was guarded and only a small inspection window was available to check the ring and roller condition on one side.

IDENTIFYING THE PROBLEM
The dryer rollers were inspected through the available inspection window. Vibration readings were not collected. On observing, it was noticed that the impacting sound recurred in every single support roller rotation. So the problem was narrowed down to the support roller assembly. It was also noticed that the sound was more from one of the two support rollers. So our concentration was on the specific support roller.
Since it recurred in one revolution and was periodic, it was concluded that the issue is not related to bearings. Bearing issues will lead to frequent random impacting, and once the bearing fails, the roller gets seized than rolling with a broken bearing roller or races. Moreover, the chances of bearing failure are less as it was heavy-duty, double row spherical roller bearings.
Recommendations were provided for the inspection and repair of the associated support roller.

FINDINGS
After proper planning, the equipment was stopped for identifying and solving the issues. On inspection, it was found that the mentioned support roller got detached from the shaft. The roller was attached on the shaft by welding and the weld has failed, making the roller free on the shaft. Continuous running with free movement of the roller on its shaft further damaged its ID (Inner Dia.) thus opening up the radial gap between roller and shaft causing the impact sounds. The impacts were produced when the worn-out roller got lifted by key in every rotation and got down due to the weight of the drum when the key passes the contact line. The roller surface was found to be worn due to prolonged running/ageing.




SOLUTION
The only solution was to replace the support rollers as the radial gaps cannot be built up. But the support roller spares were not available, which compelled us to provide temporary solutions.
Recommendations were provided to re-weld the support roller on the shaft by maintaining concentricity. The repairs were done and the rollers were reinstalled. The equipment was started and the sounds have completely vanished. The equipment ran smoothly for four months after which slight impact sound resurfaced, indicating weld joint failure.
In the next shutdown opportunity, the support rollers were replaced with a new one and the equipment is restored to a reliable condition. No abnormal impacting observed after the correction.

POINTS TO PONDER
1. Support rollers will be hardened. Any welding might induce additional stresses which might lead to failure of the shaft. So it is always preferred to buy new roller assembly in case of damage.
2. PWHT on the welded joints will remove residual stress.
3. Support roller/trunnion diameters should be checked every year to understand the wear rate so that proper planning can be done for procuring and allocating the budget.
4. Heavy-duty bearings should be cleaned and re-greased during the available opportunities.

PUMP CASE STUDY 1 - REVERSE ROTATION OF VERTICAL MULTI-STAGE BOILER FEED WATER PUMP

An interesting case came when one of the vertical multi-stage boiler feed water pumps was not developing the pressure and was having high abnormal sound from the start-up after a motor overhauling. The vibration data were collected and Direction of Rotation (DOR) was checked and found that the problem is with DOR.

How the pump can run in reverse?

There are two possible cases in which the pump can run in reverse.
  •      Backflow through pump: Usually, this happens after switching off the pump and the discharge and suction NRV/valve is passing. As the discharge head is high, water flows back through the pump making it run like a turbine. Now, the pump will be driving the motor. There is no problem in this case until the motor is started. Care should be taken so as not to start the motor while it is being driven by the pump. If started, it will add additional load causing for electrical damages in motor and/or mechanical damages in pump and shaft. But as long the rotation is not disturbed suddenly, there is no possibility of damage as the torque load on each component will be acting in the same direction.
  •     Wrong connection of motor terminals: This will create more problem than the previous case. When the driver/motor rotates in reverse direction, the torque load acts in reverse direction as compared to design, causing the failure of components like lock nuts, set screws, threaded joints and will even lead to high vibrations due to off-design flow through the pump. The common thing which can be noted when the pump is rotated in reverse direction for a short period is the loose impeller lock nut.

How to check the direction?

This is easy. Use a small plastic wire or lock tags and put it carefully on the rotating shaft surface. The direction of deflection can be noted to understand the DOR. Note that stroboscope is not a good tool for understanding the DOR. If the frequency setting is less than actual speed, you’ll find the correct DOR. If frequency setting is more than actual, you’ll find it running in the reverse direction. Or, the equipment can be turned off, and when the speed is coming down, you can understand the DOR.

Effects of reverse rotation

The pumps might contain several locking nuts/screws/bolts threaded to tighten from the natural DOR. Once it is operated in reverse, the forces on these nuts/screw/bolts will be reversed causing it to become loose. When the rotors come loose, it will lead to failure of equipment, sometimes jamming which will lead to shear failure of mechanical components or will overload and burn the windings of the motor.
In this case, the lock nut came loose and the pump seal and stage bushings got damaged due to excessive vibrations.

How to prevent failures associated to reverse rotation?
  1. Care to be taken in determining the direction of pumps. Following steps can be followed to prevent the failures associated with reverse rotation.
  2. If DOR is not marked on the pump casing, mark it during the assembly by understanding the arrangement of vanes. In some cases, by looking at the casing structure, one can determine DOR. Always cross check with the manual before marking.
  3. Check and ensure the driver DOR before coupling.
  4. Use Loctite thread lock compound (preferred 243 (blue), removable medium strength) on impeller lock nut thread to prevent it from becoming loose. 
  5. If the pump is being driven by backflow, bring the pump to rest before starting it again. Replace the passing valves as soon as possible to avoid the problems.