New Bentley Nevada ADRE 408
We are pleased to announce the latest portable test equipment upgrade for EMC. We have acquired two Bentley Nevada ADRE 408 DSPi units, each equipped with 16 channels for collecting housing vibration in the horizontal and vertical planes, as well as shaft displacement on motors and driven equipment.???????????????????????????
EMC Engineering employs advanced test equipment to perform baseline testing on equipment before removal and overhaul. These tests are repeated during the commissioning process after the equipment has been repaired, before returning to service. The graphs presented in this engineering newsletter were generated using the Bentley Nevada ADRE SXP software, which provides analysts with a comprehensive method for understanding the rotodynamic behavior of equipment and its frames' response to the rotors' excitation forces transmitted through the bearing housings.
The ADRE 408 units continuously record and collect data during start-up, operation, and shutdown phases. They facilitate the study of bode plots, making it easier to understand the natural frequencies of equipment. The setup process for this equipment is complex and can take up to a day in the field, depending on conditions and equipment accessibility, such as escort access or the badging process. Therefore, this equipment is primarily reserved for complex troubleshooting projects or critical assets undergoing major overhaul maintenance in the various industries that EMC serves.
Recently, EMC was offered an opportunity to service a motor and blower equipment train, which included a baseline performance test to comprehensively understand the rotodynamic behavior of the rotating equipment train, as well as the response of the equipment housing to the rotodynamic excitation forces. Additionally, an analysis of more intricate sources of excitation, such as electromagnetic forces in the electric motor that powers the blower, was carefully conducted before disassembling the equipment. This helped focus the Engineering forensic inspection and investigation on specific components of interest while the motor was disassembled in the shop for inspection and before disassembling the blower in the field.
This newsletter presents findings from vibration data collected with a Bentley Nevada ADRE 408 DSPi during baseline testing. Proximity probes were utilized to study the rotodynamic behavior of a coupled rotating equipment train, together with housing accelerometers, all simultaneously collected and referenced off the same key phasor. ADRE SXP software was used for post-processing the data, generating the plots presented in this newsletter.
The newsletter also features several miniature case histories from the same job EMC recently executed, including photos of equipment component damage that coincided with classical FFT signatures in Vibration Analysis training. The primary focus of this newsletter is EMC's capabilities in continuously monitoring critical assets for its customers across North America and Canada, including Commercial Nuclear Power and Municipalities, Paper Plants, Steel Mills, Large Laboratory and Health Care Facilities, Food Processing Plants, Ship Repair, Space Exploration, and Bridge Tunnel Exhaust Fans, among others.
EMC welcomes comments from Vibration Analyst experts in the industry or anyone curious about the services it offers. Please feel free to share your comments publicly or privately via LinkedIn Direct Message. EMC understands the time constraints that most Rotating Equipment Experts face and appreciates the time taken to further stimulate the interest and development of the exciting world of Vibration Analysis.
As readers peruse this Newsletter, envision the future of Vibration Analysis where Machine Learning and Artificial Intelligence can detect pattern signatures and process vibration as a variable along with other variables such as sound, temperature, current, voltage, hall sensors, pressure, flow, and more. Combining this with pictures of damaged rotating equipment, individuals who are responsible for maintaining critical assets can improve the overall reliability of their managed rotating equipment assets, even if they are not experts in Vibration Analysis. The industry now offers a staggering number of IoT options, and EMC can help guide towards the mature options that we believe would benefit the mission to improve the reliability of Electric Motors, Blowers, and Pumps the most.
For those who are stopping and checking in for this read from a motor shop, and you are an EASA member, we highly recommend visiting the EASA Emerging Technologies Committee website. This committee has done an incredible job of helping motor shops understand the available options for providing subscription based IoT monitoring of any asset, critical or not, so that you can continue to provide top decile service to your customers as well.
The image above displays a 2-pole induction motor typically running across the line, and at start. This motor is driving a blower, and the slip speed varies depending on the load, which is adjusted by variable inlet guide vanes. The motor has a voltage rating of 4,160 Volts and a full load amperage rating (FLA) of 489 amps. It is designed to deliver 4,600 horsepower to the blower.
To investigate the behavior of the motor and blower, a pair of proximity probes were installed on the opposite drive end (ODE) and drive end (DE) of the motor shaft. Additionally, another pair of proximity probes were placed on the inboard (IB) side of the blower, between the coupling train and the IB bearing housing. Unfortunately, the outboard (OB) end of the blower does not have an accessible area for the installation of proximity probes.
In addition to the proximity probes, pairs of housing accelerometers with chips to integrate the signal to velocity were placed on the ODE and DE bearing housings of the motor. Another two housing accelerometers were installed on the IB and OB ends of the blower bearing housings. The accelerometers were used to measure the vertical vibration of the blower bearing housing near the locations where the client has sensors installed to protect their equipment.
A laser tach was utilized in conjunction with reflective tape on the coupling train to activate a once-per-revolution signal that facilitated the measurement of the phase angle of vibration. All channels were consistently monitored using a Bentley Nevada ADRE 408 DSPi data collector to enable post-processing in the ADRE SXP software.
The primary objective of this setup was to collect data on the shaft displacement and housing vibration of both the motor and blower. This data can be analyzed to evaluate the rotor dynamic behavior and housing response of the entire equipment train. It can also provide a baseline for the equipment before any repair work is carried out. Furthermore, this data can assist in identifying any necessary adjustments or repairs that may be required to enhance the performance of the equipment for the customer. Ultimately, these efforts could contribute towards extending the equipment's lifespan beyond its previous limit.
Provided below is the overall trend of the DE motor shaft. As apparent, the equipment train undergoes a substantial change in vibration and magnitude as the equipment starts from cold and heats up to thermal steady state.
The miniature case histories and topics covered in this newsletter include the following:
1.?Using a Bode Plot to identify rotor critical speeds during coast down.
2.?A fun academic exercise of why one should consider high speed balancing.
3.?Looseness, due to excessive babbitt bearing running clearances.
4.?Mechanical misalignment as the dominant 2X’s frequency.
5.?Oil whirl into a rub, due to excessive bearing clearances
6.?High-resolution FFT data of a 2-pole motor, detecting rotor bar frequency
7.?Orbit signatures identifying misalignment, due to changes in hot alignment.
Topic 1:?Using a Bode Plot to Identify Rotor Critical Speeds During Coast Down
A Bode plot of the coast down below shows where EMC monitored the shaft displacement of the blower just outside of the labyrinth seal of the IB blower bearing, located next to the coupling between the coupling key and the IB blower bearing housing. Bode plots are stacked graphs representing phase and amplitude on the Y axis and displacement on the X axis.?Based on the Bode plot below, the blower rotor’s 1st critical is 2,100 RPM.?
Topic 2:?A Fun Academic Exercise of Why One Should Consider High Speed Balancing
A rotor is considered flexible when its first natural frequency (also known as “first critical”) is below its operating speed, which is commonly observed in turbines, high-energy multi-stage split case pumps with several stages, and other rotating machinery. Physicists and engineers have determined that the natural frequency of any object can be most simplistically expressed mathematically as follows:
Calculating the stiffness of a flexible rotor by pen and paper requires a solid understanding of the rotor's geometry, material properties, and boundary conditions. Here are the general steps for calculating the rotor stiffness by pen and paper:
Determine the geometry of the rotor, including its length, diameter, and cross-sectional shape.
Calculate the second moment of area of the rotor cross-section, which is a measure of the rotor's resistance to bending. For a circular cross-section, the second moment of area is as follows:
Determine the material properties of the rotor, including the Young's modulus and Poisson's ratio. The Young's modulus is a measure of the material's stiffness, while the Poisson's ratio describes the material's deformation under stress.
Apply the appropriate boundary conditions for the rotor, such as fixed or free ends.?In this case for example, the rotor is fixed and restrained on both ends and allowed to freely move axially in the journal bearings.
Use the following equation to calculate the stiffness of the rotor:
It is important to remember, Young's Modulus of Elasticity is a fundamental material property used to quantify its stiffness, also known as the spring stiffness constant. It is defined as the ratio of stress to strain under elastic deformation, indicating that the material can fully recover its original shape once the stress is removed. Young's modulus is usually measured in pressure units, such as Pascals (Pa) or pounds per square inch (psi). British scientist Thomas Young first introduced this concept in the early 19th century, and it has become a widely used analytical tool in the Vibration Analysis community. It is frequently employed to represent a shaft supported by bearings as a simplified mass-spring system to study and predict rotodynamic behavior.
With an understanding of this theory, one can see rotors with larger L/D?ratios will typically have a 1st critical below running speed for applications 1800 rpm or higher.?In other words, rotor stiffness can be simplified and expressed as the following for purposes of generally assessing if a high-speed balance of a rotor is required:
From this, it is easy to deduce that the rotor’s stiffness is proportional to the diameter of the shaft relative to the length.?I.E. the rotor is stiffer with a larger shaft diameter, which one would also expect via common sense.
For more complex rotor geometries or boundary conditions, finite element analysis (FEA) may be necessary to calculate the rotor stiffness accurately. Here are the general steps for calculating the rotor stiffness using FEA:
FEA is a powerful tool that can be used to model more complex rotor geometries and boundary conditions, as well as to account for other factors such as damping, which are not considered in the pen-and-paper method. However, it requires specialized software and expertise to use effectively.?EMC can perform this at the request of our customers, where we typically partner with Mechanical Solutions Inc to assist us on more advanced analysis for EMC’s customers.
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If the rotor were frozen at a specific moment in time, with the keyway in the same position and the rotor spinning both below and above its first critical speed, the rotor would deflect in a specific manner between and outside the supporting bearings, as depicted in the following illustration. In this scenario, the displacement of the rotor shaft outside the IB bearing would be 180° out of phase with its position when it was spinning below its first critical speed. This phase shift is well-known within the Vibration Analyst community as the "infamous 180° phase shift of the heavy and high spot," which occurs when balancing rotors operating above their first critical speed.
Balancing a blower rotor involves material removal from the shroud wall of the blower wheels using a grinder. Typically, only two planes are used for this process, but multi-plane balancing is possible, if necessary, since there are more than two blower wheels available to use as balance planes. For electric motors, pumps, and blowers with operating speeds below the rotor's first critical speed, balancing can typically be accomplished using industry standards for permissible residual imbalance, which account for the operating speed and ensure low levels of vibration when the rotor is returned to service.
However, for flexible rotors, balancing to better than 4W/N below its first critical speed may result in the material removal site on the blower wheels being located at a different position than where it needs to be during operation above the first critical speed. This may require corrections on a different balance plane entirely, or up to 180° away from the original material removal site. As a result, major balance corrections of the rotor at speeds below its first critical speed are not recommended and should be avoided if possible.
To ensure that the entire rotating equipment train is balanced as low as reasonably achievable, high-speed balancing of the rotor should be performed if possible. This will increase the likelihood of a seamless Site Acceptance Test (SAT) with minimal balance corrections required after coupling up the motor rotor to the blower rotor. Any necessary trim balance corrections in the field would be limited to correcting thermal effects of the equipment train and residual imbalance in the coupling, rather than utilizing a balance plane outside of the blower’s bearing housing to correct for imbalance of the blower’s rotor while it is operating above its first critical (i.e. applying balance weights at the coupling).
Topic 3:?FFT Signatures Identifying Looseness, Due to Excessive Babbitt Bearing Running Clearances
When there is looseness in rotating machinery, the shaft can oscillate back and forth within its bearings or housings during rotation. This can result in the shaft impacting against the bearing walls, generating a vibration signal in the time-domain that contains a series of impulses at regular intervals corresponding to the running speed.
The frequency of these impulses increases with the severity of the looseness, resulting in higher multiples of the running speed. For example, if the clearance is large enough for the shaft to oscillate back and forth twice as fast as the running speed, impulses will be generated at twice the frequency of the running speed in the time-domain. If the clearance is even larger and the shaft oscillates back and forth three times as fast as the running speed, impulses will be generated at three times the frequency of the running speed in the time-domain.
These impulses can be detected and analyzed using time-domain signal processing techniques. By identifying the specific frequencies and amplitudes of the impulses, it may be possible to diagnose the presence and severity of looseness in the rotating machinery and take corrective action to reduce the vibration and improve performance.
Below are examples of high resolution FFTs post processed from ADRE SXP for both the ODE and DE bearings, where there is clear presence of DE bearing damage and excessive clearance in both bearings.
Topic 4:?FFT signature identifying mechanical misalignment as the dominant 2X’s frequency
When the 2X running speed peak is greater than the 120Hz peak in an FFT analysis of a 2-pole induction motor, it suggests that the misalignment fault is more severe than the electrical fault. The 2X running speed component is related to the vibration caused by the rotation of the motor's shaft, while the 120Hz component is related to electrical faults such as unbalanced voltage or current, voltage fluctuations, or magnetic flux variations. Typically, a well-functioning motor will have a larger 120Hz component than the 2X running speed component.
To increase the lines of resolution in an FFT analysis, there are several options available. Increasing the sampling rate or duration of the data acquisition period can increase the total number of samples in the time-domain signal and therefore increase the frequency resolution of the FFT analysis. Thes settings are dictated by the analyst’s choice of an Fmax and lines of resolution settings in the software for most time-domain Fast Fourier Transformation processing algorithms used in the industry.
Specialized vibration analysis tools such as vibration analyzers can provide high-resolution FFT analysis. These tools typically have built-in features designed to capture and analyze high-frequency vibration signals, such as those associated with misalignment faults in 2-pole induction motors. By adjusting the frequency range, lines of resolution, and window type used in the FFT analysis, these tools can optimize the analysis for specific faults.
The FFT analysis below clearly indicates that mechanical misalignment is the dominant excitation force, surpassing the 120Hz, 2X line frequency component, and is the primary source of energy causing bearing damage. Notably, these FFT snapshots were retrieved from the ADRE SXP software of the continuously captured data during the initial startup of the machine when it was cold and at thermal steady state when the machine was hot. As the equipment train gradually warmed up, the misalignment peak at 2X’s slip speed diminished by approximately 50%, suggesting an improvement in the alignment. It is important to highlight that this observation underscores the importance of conducting vibration analysis during different operating conditions to obtain a comprehensive understanding of equipment behavior.?
This observation provides an explanation for the increased damage observed in the DE bearing as compared to the ODE bearing. The DE bearing likely experienced greater wear and tear during equipment startups, which resulted in its degraded condition. Consequently, it is recommended to minimize the frequency of equipment starts due to the potential for misalignment changes from cold to hot conditions, which could exacerbate bearing damage.
Topic 5:?Orbit Signatures Identifying Oil Whirl Progressing to a Rub
When oil whirl occurs, the rotor experiences eccentric motion due to the interaction between the rotor and the fluid film in the bearing. This eccentric motion can produce sub-synchronous vibrations that can cause the bright blank dots on an orbit plot to move around as the shaft rotates. If the bright blank dots are not overlapping with each other on each rotation of the shaft, it could indicate the presence of sub-synchronous vibrations associated with oil whirl. These vibrations are typically between 0.43 to 0.45 times the running speed of the machine.
Traditional vibration analysis techniques may not be effective in detecting sub-synchronous vibrations caused by oil whirl. Advanced techniques such as order analysis and time-synchronous averaging may be required to detect and diagnose these vibrations. Order analysis can separate the vibration signals by their frequency and determine the source of the sub-synchronous vibrations. Time-synchronous averaging can help identify the sub-synchronous vibrations by averaging the vibration signal over multiple shaft rotations.
Oil whirl is typically caused by an unstable fluid film in the bearing that supports the rotating shaft. The fluid film can become unstable due to various factors, including changes in operating conditions, bearing design or clearance, viscosity of the lubricant, and rotor dynamics. When the fluid film becomes unstable, it can cause the rotor to experience an eccentric motion, resulting in oil whirl.
Excessive clearance in bearings can be caused by wear, corrosion, and poor maintenance practices. This can contribute to the occurrence of oil whirl by reducing the stiffness of the bearing and causing the rotor to experience an eccentric motion. Proper maintenance and monitoring of the machine can help to prevent oil whirl and detect it early if it occurs, reducing the risk of equipment damage and downtime. Additionally, oil whirl can also be caused by misalignment, unbalance, or other mechanical issues that affect the dynamic behavior of the machine. High-speed rotation, high loads, and uneven wear on the bearing surfaces can also contribute to the occurrence of oil whirl.
Based on the below FFT plots, it is evident that the Blower shaft was exhibiting whirl in the inboard (IB) bearing during startup, which progressed to a rub as the machine warmed up and the oil viscosity reduced. The initial sub-synchronous frequency observed after startup indicated the possibility of oil whirl at 0.45X rather than a rub, and then eventually transitioned to a dominant peak at 0.5X running speed, which is indicative of a rub.
Below is a picture showing the rub damage identified of the blower bearing after equipment tear down and disassembly:
Topic 6:?Rotor Bar Pass Frequency Identifies Rotor Issues
Vibration accelerometers installed on the motor's bearing housing can detect a dominant frequency that corresponds to the Rotor Bar Pass Frequency (RBPF), along with a 120Hz sidebands.
The rotor bar pass frequency generates a frequency that matches these calculations through a phenomenon called bar passing. As the rotor bars pass the rotor core slots during rotation, they produce a magnetic field that interacts with the stator windings, causing a vibration. This vibration produces a frequency that is proportional to the number of rotor bars and the running slip speed of the motor. In the case of the given motor, there are 40 rotor bars that produce a frequency that matches a calculated value of 143,560 CPM. Hence, a presence of a dominant frequency at 2392.667 Hz corresponds to the Rotor Bar Pass Frequency (RBPF) detected by the vibration accelerometers can be explained by the bar passing phenomenon.
Based on these identified issues, it was recommended to address the root cause by rebuilding the rotor entirely. In the case of rebuilding the rotor, it is essential to ensure that the dimensions and tolerances of the rotor bars and rotor core slots are within the manufacturer's specifications or tighter. Alternatively, temporary repairs may include tightening the rotor bars in the core slots by means of swaging to eliminate the excessive clearance and prevent further migration of the rotor cage. The preferred approach will depend on the severity of the damage and the extent of the necessary repairs. It is important to note that early detection and prompt resolution of such issues can prevent further damage to the motor and ensure its optimal performance.
When diagnosing rotor bar pass frequency issues on high-speed, two-pole motors operating at 3,600 RPM, it is crucial to utilize a data acquisition system that is capable of adequate lines of resolution and fmax. This is necessary to ensure the accurate detection of the high-frequency Rotor Bar Pass Frequency (RBPF) and the precise Knife Edging of its peak in the Fast Fourier Transform (FFT) analysis. Moreover, an adequate number of lines of resolution in the FFT enables the study of the sideband frequencies that are associated with the RBPF, providing a comprehensive understanding of the issue at hand.
The presented FFT indicates the existence of excessive clearance between the rotor bars and slots, which leads to the vibration of the bars at the excitatory frequency of 120Hz generated by the 60Hz power supply. Consequently, axial migration of the rotor cage towards the left direction occurred over time, resulting in an increase in vibration levels and a reduction in motor performance. Additionally, during the motor start-up process, there was excessive rattling, which helped to clarify the root cause of the problem.
The dominant peak in the FFT analysis below corresponds to a frequency of 2392.667 Hz, which equals 143,560 cycles per minute (CPM). This frequency is derived from the 40 rotor bars in the motor, where the running slip speed of 3,589 RPM multiplied by 40 results in a frequency of 143,560 CPM.
Topic 7:?Orbit Signatures Identifying Misalignment Due to Changes in Hot Alignment
In a motor to pump or blower equipment train, misalignment between the two components can have a significant impact on system operation. One area where this impact is particularly noticeable is in the shaft displacement orbit.
When proper alignment is achieved between the motor and driven equipment, the shaft displacement orbit exhibits a circular shape, with slight elliptical characteristics in the case of horizontal applications. The stability of the orbit is indicative of a smooth transmission of rotational force between the two components. In contrast, misalignment leads to an increase in the degree of flattening of the orbit, reflecting the lateral forces generated as a result. In situations of severe misalignment, the orbit may assume a figure-eight shape, signaling a significant level of misalignment that results in unpredictable movements of the rotating components.
To monitor the shaft displacement orbit and detect misalignment, shaft proximity probes are commonly used. These probes accurately measure the shaft's position and detect deviations from the expected circular orbit. They are particularly useful when monitoring the coupling train on driver and driven equipment since misalignment can occur in any part of the train.
The shaft displacement orbit is a crucial indicator of the alignment between the motor and pump in a system. Shaft proximity probes are essential to obtain accurate measurements of shaft displacement and identify any misalignment that may cause unpredictable paths for the rotating components. This is especially important for complex coupling trains where misalignment is more likely to occur.
Severe misalignment generates lateral forces that cause the rotating components to move in unpredictable paths, resulting in a figure-eight pattern in the shaft displacement orbit. The size and shape of the figure-eight pattern provide an indication of the severity of the misalignment, with a more pronounced figure-eight indicating a higher degree of lateral movement between the motor and pump.
The use of shaft proximity probes is critical in preventing equipment damage and ensuring optimal system performance by accurately capturing shaft displacement and detecting misalignment. Proper alignment between the motor and pump is, therefore, crucial for efficient and reliable system operation.
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