May 2016

Condenser Monitoring 10 • Steam Turbine Rotor 14 • ASME: Wave Energy Conversion 18

ENERGY-TECH A WoodwardBizMedia Publication

MAY 2016

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Printed in the U.S.A. Group Publisher Karen Ruden – kruden@WoodwardBizMedia.com General Manager Randy Rodgers – randy.rodgers@Woodwardbizmedia.com Managing Editor Kathy Regan – editorial@WoodwardBizMedia.com Editorial Board (editorial@WoodwardBizMedia.com) Bill Moore – Director, Technical Service, National Electric Coil Ram Madugula – Executive Vice President, Power Engineers Collaborative, LLC Kuda Mutama – Engineering Manager, TS Power Plant Tina Toburen – T2ES Inc. Editorial views expressed within do not necessarily reflect those of Energy-Tech magazine or WoodwardBizMedia.

FEATURES

6

New ultrasonic flow meter technology for detection of condensate in HRSGs

By Bill Carson, Electric Power Research Institute

COLUMNS

10

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Condenser monitoring instrumentation identifies abnormal condition

By Collin J. Eckel, Intek, Inc. and Barry T. Brown, Sr. Plant Engineer for East Kentucky Power Cooperative

14

Turbine Tech

The root cause analysis challenges of steam turbine rotor material loss

By Matt Scoffone, consulting engineer for TG Advisers, Inc.

25

Machine Doctor

Gear tooth cavitation erosion

By Patrick J. Smith

ASME FEATURE

18

Wave energy conversion: control of the buoy heave motion

By Ossama Abdelkhalik, Shangyan Zou and Rush Robinett; Michigan Tech University and Giorgio Bacelli and David Wilson; Sandia National Labs

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Maintenance Matters

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May 2016

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EDITOR’S NOTE

What’s new?

New format, new webinars, new live symposium, new editor Change can be scary. Change can also be rewarding and open doors to new opportunities. I’m more of a glass half-full type person and always look forward to something new.

Welcome to Energy-Tech’s second quarter issue of 2016 and my first as the new editor! I’m quickly getting acquainted with turbines, emissions controls and heat exchangers which I’m finding is not too far off from my years in the scientific equipment industry. I’m enjoying learning about the power generation industry and look forward to meeting the experts behind many of the articles that we present to you, our readers, throughout the year.

As I gathered the articles for this issue, I was pleased with the content and believe that we’ve put together a good selection of articles for you this quarter. We hope you are enjoying our new editorial format and are receiving the weekly email newsletters that will keep you informed between our print issues, which are mailed in February, May, August and November.

This year is bringing lots of changes to Energy-Tech, not only with how we publish the magazine and my role as the new editor, but also in what we plan to offer. We have a great mix of online webinars offered through Energy-Tech University and are partnering with Environment One Corp. to present a live Generator Auxiliary Systems Symposium designed to increase awareness of critical auxiliary systems for hydrogen cooled generators, with particular focus on gas cooling systems, associated safety standards, operating procedures and risk mitigation technologies.

CALENDAR May 10 – 12, 2016 Diagnosing and Correcting Gas and Steam Turbine Vibrations www.Energy-Tech.com/Vibration June 13-14, 2016 Fugitive Emissions Summit Americas Houston, Texas www.fugitive-emissions-summit.com June 26-30, 2016 ASME 2016 Power & Energy Conference & Exhibition Charlotte, NC www.asme.org/events/power-energy August 1-3, 2016 Generator Auxiliary Systems Symposium Schenectady, NY www.Energy-Tech.com/Gen-Sym Dec. 13-15, 2016 Power-Gen International Orlando, Fla. www.power-gen.com Submit your events by emailing editorial@woodwardbizmedia.com.

Find out more about both at www.energy-tech.com.

And, finally – are you going to the 2016 ASME Power & Energy Conference in Charlotte, NC? This conference brings together all of ASME’s world class conferences on Power Generation, Energy Sources and Energy Sustainability and is always a great learning and networking experience. General Manager Randy Rodgers will be attending the conference as I celebrate the wedding of my youngest son. We welcome your ideas for future issues of the magazine. Stop by booth #317 to meet Randy or email us at editorial@woodwardbizmedia.com.

Thanks for reading.

Kathy Regan

4 ENERGY-TECH.com

May 2016

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FEATURES

New ultrasonic flow meter technology for detection of condensate in HRSGs

By Bill Carson, Electric Power Research Institute

During startup of many heat recovery steam generators (HRSGs), inadequate or poorly designed drain systems are unable to detect and drain condensate in the high-pressure superheaters (HPSHs) and reheaters (RHs). Ineffective drainage of these tubes is a leading cause of premature failure of HRSG pressure components.

Many of the HRSGs currently in service have ineffective drain systems; even those recently installed with automatic drains intended to detect and effectively drain condensate during startup are incapable of doing so under all startup conditions.

The Electric Power Research Institute (EPRI) has been working with Competitive Power Resources Corp (CPR) of Palmetto, Florida to study the potential of an ultrasonic technology to detect, for the first time, the presence of condensate in the headers and tubes of HRSGs under all startup conditions. The ultrasonic equipment, a modified version of a stateof-the-art flow meter manufactured by Flexim Americas, would allow operators to automatically drain the condensate and prevent or minimize major equipment damage.

cent tubes. These temperature differences, in turn, cause excessive stress at tube-to-header welds as a small portion of tubes are subjected to rapid and severe quench cooling, while the remainder of the tubes attached to the upper and lower headers remain close to gas temperature.

The result can be severe thermal-mechanical fatigue to headers, tubes, steam piping and tube-to-header joints (Figure 1). This damage accumulates over time and often goes undetected until a crack initiates and propagates through the tube, header or pipe wall (Figure 2). Repair/replacement of tubes and headers caused by this failure mechanism can be very costly and result in significant lost generation associated with the unplanned outage time.

This failure mechanism has been a significant and longtime industry problem, resulting in avoidable deterioration of unit reliability and unnecessary maintenance costs. The problem is mainly an issue in horizontal-gas-path HRSGs but can also affect vertical-gas-path HRSGs.

To date, the technology has been installed and tested in three plants.

Additional field testing is planned at other host plants during 2016.

Detecting the presence of condensate At startup of combined-cycle natural gas plants, condensation occurs in the high-pressure SHs and RHs of the HRSG. If this condensate is not drained, it remains within the headers, pipes and tubes at the same time that steam flow is initiated through them. This interaction creates a pressure drop that forces the undrained condensate through the downstream sections of the SHs and RHs. When the condensate reaches the significantly hotter SH and RH components, it causes significant temperature differences in headers and between adja6 ENERGY-TECH.com

Figure 1. Thermal-mechanical fatigue crack in SH tube from repeated quenching.

May 2016

FEATURES inside the pipe using the transit-time difference method. It exploits the fact that the transmission speed of an ultrasonic signal depends on the flow velocity of the carrier medium. An ultrasonic signal moves slower in water than in the steel pipe wall.

Signals generated by the Flexim equipment enable the plant’s distributed control system (DCS) to automatically open each drain valve only when it must be opened to discharge water, thereby minimizing loss of drum pressure and live steam. In many cases, use of the liquid detector avoids the cost of significant drainpipe modifications by facilitating drainage of non-optimum drainpipes.

This ensures that all water is removed before steam flow in the HRSG begins and avoids the tube damage that otherwise results from inadequate draining.

Figure 2.

Moreover, the problem is likely to become a much greater issue as the HRSG fleet experiences increasing demands for cycling operations and plants are cycled on and off load more frequently, expending at each startup the limited fatigue life of tubes at their header attachment welds.

In the past, one approach to this issue involved retrofitting large, instrumented chambers called “drain pots,” which could detect water in the drain system and generate control signals that could be used to open and close the drain valve at the appropriate times. Unfortunately most operators found the retrofit of drain pots impractical due to their large size and high cost.

What is needed is a smaller, reliable, less expensive way to know when water is in the drainpipe and when steam is in the drainpipe so that automatic controls can open the valve to let the water out and close the valve to keep the steam in.

Presently most HRSG operators have no way to determine when drains need to be open and when they may have to be closed. Premature closure of the drains can result in severe damage to piping, tubes and headers, while excessive use of the drain system wastes thermal energy, resulting in undesirable decay in drum pressure, and overheating of drain components.

Ultrasonic flow meter technology The Flexim ultrasonic liquid detection technology has the potential to detect water versus steam in HRSG drainpipes. The technology, which is clamped to the exterior of HRSG drain piping, uses two ultrasonic transducers to interrogate the interior of the drainpipe (Figure 3). The detector takes measurements

Because the Flexim technology requires no penetration of the drainpipe’s pressure boundary, it removes a potential safety concern associated with use of drain pots that employ level switches or capacitance-type level probes for feedback control of the desired water content in the drainpipe system. Also, no part of the Flexim sensor is in contact with water or steam to become contaminated, no welding is required, and no vertical elevation is sacrificed for installation of the sensor.

Applications at two HRSGs To test the operation of the Flexim flow meter in a power plant environment, the technology was installed at three plants: Oglethorpe Power Company’s (OPC’s) Thomas A Smith Energy Facility, a 1,250-MW plant near Dalton, Georgia; Public Service Electric & Gas (PSE&G’s) Bethlehem Energy Center, a 757-MW plant in Albany, New York and We Energy’s (WEE’s) Port Washington plant in Port Washington, Wisconsin.

The HRSG at OPC’s T.A. Smith Energy Facility is experiencing ~150 starts per year, which is well outside the design criteria of 10 starts per year. These increased starts were leading to the formation of condensation migration, which was causing fatigue-induced failures.

The field trial of the flow meter technology project demonstrated that the flow meter technology can distinguish the difference between saturated liquid and vapor within the drain system during both a cold and hot startup. EPRI and OPC worked with the vendor to incorporate equipment modifications and improvements and to enhance installation techniques.

May 2016 ENERGY-TECH.com

7

FEATURES

Figure 3. Flexim ultrasonic liquid detection system installed on a 4-in. drainpipe

The first permanently mounted liquid detection system was installed and tested at one HRSG at the T.A. Smith facility in 2014. Following additional testing by EPRI, resolution of several technical issues, and upgrading of hardware and software, liquid detection systems were installed in 2015 on three other HRSGs at the facility. Operating parameters and procedures have been revised to better integrate drain control optimization during unit startup.

PSE&G’s Bethlehem Energy Center provided a HRSG host site with a different size, configuration, and operating mode compared to the OPC sites. The Bethlehem test site, thereby, involved important differences for installing the flow meter system, which, in turn, caused a redesign of both the installation hardware and detection software. Without this host site, the need for this redesign would not have been known, and could have caused a setback in the success of the expected industrywide installations. Thus the smaller, tighter configuration of the Albany plant is anticipated to play a crucial role in bringing this state-of-the-art technology to the HRSG fleet.

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WEE’s Port Washington plant provided yet another host HRSG with different drainpipe size / arrangements. The field tests there involved simultaneous installation of multiple liquid detection systems on different pipe sizes as well as confirming the performance of additional methods of interface with the plant’s DCS. Based on field tests at these three host facilities, the EPRI/ CPR team has developed application-specific drain valve control logic for use with the Flexim liquid detection system.

Following confirmation of the performance of this logic during additional field-testing, planned for 2016, the system is expected to be available for commercial application by Flexim. ■ Bill Carson manages EPRI’s Combined Cycle HRSG and Balance of Plant Program. He joined EPRI in 2008, after working in various capacities at Dynegy for 18 years. Carson has more than 30 years’ experience managing the construction and ensuring the safe and reliable operation of plants in the power industry. He can be reached at editorial@woodwardbizmedia.com

May 2016

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MAINTENANCE MATTERS

Condenser monitoring instrumentation identifies abnormal condition

By Collin J. Eckel, Intek, Inc. and Barry T. Brown, Sr. Plant Engineer for East Kentucky Power Cooperative

Upset operating conditions that degrade condenser performance have a negative impact on cycle efficiency, causing increases in fuel consumption to meet power demands. These inefficient abnormal operating conditions can occur rapidly and often go undetected and/or require extensive time and resources to investigate and resolve. Standard condenser instrumentation typically provides little to no early detection or guidance for troubleshooting and determination of the root-cause for the abnormal operating condition.

The following case study shows how an investment in condenser instrumentation can economically result in improved early detection, trouble shooting, and correction of upset conditions. This increased visibility in condenser performance and operation will help to maintain plant efficiency.

Unit description and background A condenser monitoring system, using test-grade instrumentation, was installed on a 340 MW coal-fired unit at East Kentucky Power Cooperative’s Spurlock Station, with the goal of continuously monitoring condenser performance. The condenser is a Westinghouse single pass single shell design, with divided waterboxes supplying circulating cooling water to two tube bundles from a mechanical draft cooling tower. The east side tube bundle is fed from a branch tee, off the main circulating cooling water supply line, that tees upward into the east inlet waterbox. The straight-through side of the tee continues further and elbows upward into the west inlet waterbox. [This tee and elbow configuration plays an important role in the findings]

Figure 1.

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May 2016

MAINTENANCE MATTERS Previous analysis of the existing plant instrumentation had revealed uncertainties in the condenser pressure, inlet circulating water temperature, and outlet circulating water temperature values. There was no existing instrumentation to directly measure circulating water flow rates nor means to continuously measure condenser air in-leakage. To address these issues new instrumentation was installed, in the form of a condenser monitoring system, which comprised the following instrumentation: • RTDs (Qty. 2): Measure circulating water temperature into each inlet waterbox.

• RTDs (Qty. 8): Measure circulating water temperature out of each outlet waterbox. Four RTDs in each outlet circulating water pipe to account for thermal stratification.

• Combination Pressure-Temperature probes (Qty. 2): Measure condenser shell side steam pressure and temperature. One P-T probe located at the circulating cooling water inlet end of the condenser, and the other probe located at the circulating cooling water outlet

end of the condenser, on the adjacent tube bundle. Both probes are located 2 ft. above the top of the tube bundle.

• Differential Pressure (DP) meters (Qty. 2): Measure circulating water flow through each flow path (A flow path is defined as the flow from an inlet waterbox to an outlet waterbox). Each DP meter is installed across the respective outlet waterbox and the circulating cooling water pipe transition.

• RheoVac Air In-Leak Monitors (Qty. 2): Located at air off-take for each bundle, and on the common air offtake line. RheoVac probes can be moved to different locations as needed.

The DP meters were field calibrated using a pitot traverse method, at the cooling tower risers, for three flow rates; two circulating water pumps running, one circulating water pump running, and a mid-flow rate with both circulating water pumps running with discharge valves throttled. The cooling tower blowdown and cooling water supply to plant equipment were isolated prior to DP meter calibration.

Figure 2.

May 2016 ENERGY-TECH.com

11

MAINTENANCE MATTERS

Figure 3.

Data from all of the instrumentation is logged in a main processing unit that calculates and graphically displays the desired information.

For this particular installation steam pressure, steam temperature, circulating water flow inlet and outlet temperature, air in-leakage, and circulating water flow rate are measured and condenser duty and cleanliness factor are calculated in the main processing unit.

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Event details The event described in this report occurred as the unit was coming out of an outage. The condenser monitoring system installation and calibration was completed shortly after the unit went online; however, due to maintenance work, only one circulating water pump was initially available for service. The second circulating water pump went online 11/22/15 Actions taken At a later date, while the unit was offline, the inlet waterboxes were inspected for debris. A significant amount of foreign debris was found lodged in the tubes and lying in the bottom of the west inlet waterbox. Further inspection revealed that additional debris had accumulated in the circulating cooling water supply line (downstream of the previously described tee) that feeds the west inlet waterbox. Approximately ten cubic feet of debris was removed from the water box, circulating cooling water supply line, and circulating water pump suction basin.

Pictures from the waterbox and piping inspection are shown in figure 2. Efforts to remove the debris were time constrained by operational requirements, so not all debris was removed from the tubes in the west tube bundle at that time. The services of a condenser cleaning contractor will be required to fully remove the debris from the condenser tubes at the next opportunity.

Results The startup of the unit following the macrofouling event is shown in the right half of figure 1. The total flow rate increased to ≈157 kGPM. The condenser cleanliness factor increased to 70% even though some debris remained embedded in the west bundle’s tubes.

May 2016

MAINTENANCE MATTERS Conclusions Upon further investigation, the source of the debris was found to be the repair of an underground cooling tower makeup water line that occurred adjacent to the online circulating water pump basin between 11/03/15 and 11/05/15, shortly after the initial unit startup. The fouling occurred in the west inlet waterbox due to the tee and elbow configuration of the circulating water supply line; the trajectory of the entrained debris carried it past the branch tee of the east waterbox to the end of the line at the west waterbox. This fouling event occurred shortly after the initial startup of the condenser monitoring system and rapidly caused a significant decrease in condenser performance. This type of abnormal operating condition was not obvious on the less sensitive plant instrumentation that simply measured DP across circulating cooling water supply and return lines. While the plant’s DP measurement has a slight upward trend, it would have been difficult for operators to detect an abnormal operating condition. As shown in Figure 3, the individual path flow rates [light blue and dark blue] indicate an abnormal condition much more clearly than the plant instrument does [orange].

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m If Spurlock Station had not installed the condenser monitoring system, it is likely the unit would have run for an extended period of time with higher than necessary turbine backpressure at the expense of increased fuel consumption.

While it is not possible to estimate the potential cost of this macrofouling event, it is easy to state that early detection of macrofouling will help Spurlock Station achieve East Kentucky Power Cooperative’s goal of generating electricity at the lowest possible cost for its member owners. This specific example demonstrates how an initial investment in a condenser monitoring system could provide an immediate pay back for plant operators and owners. ■ Mr. Collin J. Eckel holds a B.S. in Mechanical Engineering from The Ohio State University. At Intek he has been actively engaged in the analysis of plant and Intek instrument data, supporting customers and designing diagnostic software for RheoVac users. Collin contributes to Intek’s continuing efforts to improve instrument performance by refining our calibration and instrument preparation methods, and developing software improvements. He has visited numerous power plants, assisting customers with troubleshooting and participating with other Intek staff in the inspection and evaluation of condensers. He is a member of the Intek team conducting the installation, monitoring and data analysis of our advanced condenser monitoring systems. He can be reached at editorial@woodwardbizmedia.com Mr. Barry T. Brown holds a B.S. in Mechanical Engineering from Iowa State University of Science and Technology, 2011 and B.A. in History and Geography from Northwest Missouri State University, 1990. He currently is a Sr. Plant Engineer for East Kentucky Power Cooperative at Spurlock Station, Maysville, KY, where he acts as the system expert for condensers, cooling towers, air heaters, and pumps. Prior to pursuing an engineering degree at Iowa State University, Mr. Brown worked at a coal fired cogeneration facility for a large agricultural processor as a control room operator. He has been involved in coal fired power generation and engineering for fifteen years. He can be reached at editorial@woodwardbizmedia.com

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TURBINE TECH

The root cause analysis challenges of steam turbine rotor material loss

By Matt Scoffone, consulting engineer for TG Advisers, Inc.

While working through issues on an HP/LP turbine rotor, the discussion turned to the mechanisms most often behind material loss on steam turbine rotor discs. There are a number of possible causes for material loss on a steam turbine rotor, and being able to successfully diagnose the underlying root cause is key in preventing potentially costly weld repairs. From TG Advisers experience, two of the most common causes of material loss on steam turbine rotor discs are erosion and pitting corrosion. Beyond simply determining the root cause of any material loss discovered during an outage, analyses must be carried out to verify whether repairs are necessary.

First steps: Are repairs required?

The most serious implication of rotor material loss stems from the reduction in stress-bearing area at the steeples/dovetails.

They are the primary load bearing areas on a turbine rotor disc, and they retain the blades. If enough material is removed from these highly stressed features, there is a likelihood of catastrophic failure leading to blade liberation, or worse. OEM-level analyses are necessary to determine the capability of the rotor disc geometry while accounting for the reduced area. Often, the damage is not discovered until a unit is opened during an outage, when an answer to whether repairs must be carried out is needed as quickly as possible. During the new design process, Finite Element Analysis (FEA) programs such as ANSYS or ABAQUS would be used by OEMs in defining the geometry for the rotor discs. However, these analyses are highly involved and time consuming. Compounding the issue, solid models do not always exist for legacy turbine rotor designs, further adding to the time necessary to carry out FEA. Fundamental stress calculations can be worked as an alternative method, using a few simple geometrical inputs.

Bending stress, shear stress, bearing stress, and direct stress in the rotor disc due to the pull load from the blades all need to be determined and evaluated, using a combined stress theory of failure. One of the more well-known and extensively used combined stress theories is the distortion energy theory, more commonly known as the von Mises theory.

Comparing the combined stress against a temperature-corrected yield strength Figure 1. for the rotor/disc material gives the safe14 ENERGY-TECH.com

ty factor in the as found condition which can be used as a guide for remaining disc capability considering the loss of material.

Fatigue is the other major concern when steam turbine rotor discs lose material, especially Low Cycle Fatigue (LCF) from unit stop start cycles. Non-destructive evaluation (NDE) inspections are an important step in minimizing any risk to the discs through LCF. However, NDE inspections are not perfect, and the most commonly used inspection methods of Magnetic Particle Inspection, Fluorescent Penetrant and UT Phased Array Inspections do have a minimum crack size that they can detect.

As a result, indications may be present in the rotor discs, even if the NDE report comes back clean. Evaluating the rotor disc LCF life using the combined stress previously calculated in the safety factor analysis is another important step in minimizing risk of future failure when assessing whether repairs are required.

Assuming an initial flaw size equal to the largest detected indication found through NDE, or using an assumed minimum detectable flaw size, remaining life in cycles can be evaluated.

The result obtained is the predicted number of unit start stop cycles until the flaw grows to a critical size, at which point failure can occur. Similar to combined stress theories, there are a number of formulas available for evaluating LCF, though Paris’ Law is one of the more commonly used methods. Coupled with the factor of safety, a judgement call based on real data can be made for whether repairs are necessary, or if the unit can

May 2016

TURBINE TECH

Figure 2.

continue to run to the next major inspection interval with low risk of a forced outage or failure occurring prior to that date.

Diagnosing diseases: Erosion or corrosion?

There are two main types of erosion that typically occur in steam turbines: water droplet erosion, and solid particle erosion. For water droplet erosion to occur, water droplets have to be able to form around the stages being investigated. The most common source of water droplets in a steam turbine is the steam itself; for this to happen however, the steam conditions have to allow for condensation. The Wilson line defines the area where steam transitions from superheated to saturated, at which point water droplets often form. The Wilson line is most often aligned within the latter stages of a utility sized LP rotor, or the mid stages of smaller HP/LP industrial size rotors. For water droplets to be the root cause of material loss ahead of this area, the water droplets must stem from some other feature within the system. In rare occasions, extraction lines are one such possible feature that could result in water ingress ahead of the Wilson line. Confirming that these lines are adequately drained can rule out any extraction issues, and potentially rule out water droplet erosion if the stages being investigated are upstream of the Wilson line. Water droplet erosion typically produces jagged peaks, similar to what is often seen on last stage blade tips as seen in the example photo in Figure 1. Water droplet erosion is rare on steam turbine discs, and is most often found on long LP blades near the tips. The tips of these long LP blades travel at a high velocity, as the distance that a later stage LP blade tip must travel in each rotation is significantly larger than the distance a point on the rotor body must travel.

Solid particle erosion occurs when solid particles are introduced into the steam path at a high velocity. There are numerous sources of solid particles, with one of the more prevalent being rust particles from the boiler. Solid particle velocities are highest in the first few stages of an HP turbine or IP reheat turbine, and their energy rapidly dissipates as they travel further downstream. Solid particle erosion typically May 2016

Figure 3.

occurs during startup, when the steam entering the unit is being controlled by valves that are open only a small amount.

The valves act as nozzles, increasing the velocity of the flow entering the turbine. Due to the high velocities necessary for solid particle erosion to occur, any material loss downstream of the first few stages of an HP turbine or IP reheat turbine is most likely not from solid particle erosion. Metal loss due to solid particle erosion typically yields fairly uniform surfaces absent of pitting, as shown on the diaphragm airfoils in the example photo in Figure 2.

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TURBINE TECH Oxygen pitting corrosion is another cause of material loss within a steam turbine, especially if the material damage is confined primarily to the rotor. Metal that has undergone pitting corrosion has a rough, non-uniform surface finish, with numerous circular concave pits of various depths randomly distributed across the surface, as seen in the example photo in Figure 3. The combination of oxygen and moisture often found in steam turbines during lay-up, or when shut down, forms the environment necessary for pitting corrosion to occur. Pitting corrosion is a static process, and should not occur while the unit is operating.

Any sodium chloride or salts that are present, either in a wet or a vapor form, will further accelerate the corrosion process, pos-

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sibly to significant levels. Samples of deposits that may be found on the blades, vanes, or rotor discs should be checked for the presence of salts. Turbine rotors and discs are highly susceptible to corrosion due to the low chrome content of these components, typically 1-3%. Blades and vanes usually have a much higher chrome content, 11% or more, which is enough to prevent significant corrosion from occurring.

The possibility of pitting corrosion occurring on the turbine rotor or discs can be minimized through a few steps.

Limiting the exposure of the unit to oxygen during lay-up and shut down by maintaining vacuum during short term shut downs will slow the corrosion process. However, maintaining vacuum is not always a practical solution, especially for longer outages. Another option to limit oxygen exposure is to use nitrogen blanketing in lay-up procedures when feasible. This can be a costly and difficult process to implement however. The more practical approach to minimize corrosion is to lay-up the turbine in a dry, low humidity environment. Injecting dehumidified air into the turbine is one method to accomplish low humidity, and spin cooling the unit while it is still warm can minimize any condensation that could form on the rotor discs.

Long term, ensuring proper steam chemistry by maintaining the OEM recommended chemistry guidelines will reduce the aggressive contaminants that are part of the root cause for corrosion.

Final steps and future work Any material loss should be reevaluated at A-T Controls has the products, knowledge, and future outages and inspection intervals, with the capability to handle your complete project from ¼” Ball Valves, to 48” Butterfly Valves and stress and LCF analyses updated to reflect addicomplete Automation products and services. tional damage that may be present. Independent Choose A-T Controls for your next project! consultants are always available to provide an outside opinion, or evaluate any material loss present in the system to determine the root cause of the issues. ■ Matt Scoffone is a consulting engineer for TG Advisers, Inc., Matt and the TGA team provide engineering services to gas and steam turbine and generator users worldwide. Matt is a graduate of General Electric’s Edison Engineering Development Program and holds a BSME from Rensselaer Polytechnic Institute and a MSME from Georgia Tech. Matt can be reached at editorial@woodwardbizmedia.com

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May 2016

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ASME FEATURE

Wave energy conversion: control of the buoy heave motion

By Ossama Abdelkhalik, Shangyan Zou and Rush Robinett; Michigan Tech University and Giorgio Bacelli and David Wilson; Sandia National Labs

ABSTRACT This article discusses a new method for optimizing the control of a heave wave energy converter aiming at maximizing the energy extracted from ocean waves. The recently developed singular arc control for wave energy conversion is implemented.

Excitation force estimation is a necessary step in implementing the singular arc control. A fast Fourier transform approach is adopted to estimate the excitation force. Numerical results are presented that demonstrate the efficiency of this proposed system. This effort is coupled with the ongoing advanced controls project at Sandia National Labs (SNL) to study the improvement of wave energy converters performance through the control system design. This article presents numerical experiments on the SNL’s experimental buoy.

Introduction Buoys are being used widely for sensing and collecting data. For instance moored buoys are being used for weather sensing in the sea. Several buoys are deployed in the coastal and offshore waters, e.g. in the western Atlantic, the Pacific Ocean, the Bering Sea, and the South Pacific [1]. The buoy provides the necessary power for the sensors through using batteries or solar panels, depending on the sensors being used. Maintenance is a challenge when using solar panels.

Buoys can be a source of power if they are designed and controlled for that purpose. The amount of power available in waves is huge and dense which means it can be an additional source of renewable energy, besides other renewable sources such as solar and wind energy. There are challenges that hinder the utilization of buoys for converting wave energy into electricity. One challenge is related to the power conversion mechanisms which efficiencies are not high enough to enable economic deployment of wave energy converters (WECs). Another challenge is the harsh environment in the ocean which may damage buoys if not designed to survive such harsh environments. On the other hand, one would design and control buoys such that they extract as much energy as possible from the wave; this requirement of maximizing the extracted energy may not align with the requirement of surviving the harsh ocean environment. A trade-off between these requirements is then necessary. For efficient buoy operation, it is important to solve the buoy controls problem in an optimal sense. To do that, a deep understanding for the buoy hydrodynamic interaction with waves is vital, and elegant control approaches need to be implemented.

18 ENERGY-TECH.com

This article presents one of the recent advances in control algorithms for maximizing wave energy conversion through WECs. To highlight the challenges in controlling WECs, the next section presents some details on the dynamics of WECs.

Dynamics of a heave WEC Most WEC dynamic models in the literature assume linear hydrodynamic device models. Some references, however, use linear models with uncertain dynamics to account for nonlinear effects and modeling errors [2]. The use of linear models is suitable for small buoy motions compared to the wave elevation, which is assumed to be the case in the study presented in this article. A buoy in the ocean is subject to forces that can be categorized as follows. The hydrostatic force, ƒs, is the difference between the gravity and buoyancy force. It has a spring-like effect on the buoy. If the displacement of the float from the sea surface is x, then the hydrostatic force can be written as ƒs = −k x.

The wave flow has a potential field that dictates the pressure at each point in the field. The interaction of the buoy surface with this potential field results in a force on the buoy; this force is called the excitation force, ƒe. Moreover, when a buoy moves in water, it generates waves and these generated waves create their potential field that results in additional force on the buoy. This force is called the radiation force, ƒr [3].

To control the buoy motion, we need to apply a force and we call it control force, u, defined positive in the opposite direction of increasing x. The motion dynamics of a floater of mass m can then be described using Newton’s second law of motion as: m¨x = ƒe+ƒr + ƒs−u, [4].

The control problem is to compute the history of the control force u(t) over the interval [0 T] such that the converted energy is maximized. As can be seen from the above discussion, the radiation force on the buoy is mainly a function of the buoy motion and hence it can be computed in real time and used to compute the optimal control force u. The hydrostatic force is also a function of the buoy position and hence it can be computed and used to compute u. The excitation force, on the other hand, is a function of the buoy motion as well as the wave potential field. Meaning that one needs to know the wave and its potential field in order to compute the excitation force so that one can use it to compute the optimal control force u(t).

To estimate the excitation force, physical sensors are needed to collect measurements. Typically, buoy position is measured. The buoy position, however, is result of the interaction of the wave with the buoy body and hence it is not a direct measurement ASME Power Division Special Section | May 2016

ASME FEATURE of the excitation force. Sensing the pressure at few points on the buoy surface provides measurements that are more direct to the excitation force. Pressure transducers can be used to measure pressures at certain locations on buoy surface.

To complete the optimization of the control force, one needs to use the measurements to estimate the excitation force.

This can be done in few different ways. This article presents a method that implements a Fast Fourier Transform (FFT) approach extract estimates about the wave different frequencies, their amplitudes, and phases and then use these estimated states to compute an estimate for the excitation force. Once the excitation force is estimated, we can compute the control force in an optimal sense. Recently, a rigorous development for the optimal control was developed within the context of the optimal control theory; this controller switches between a bang-bang control and a singular-Arc control depending on a switching surface. This controller is presented in detail in reference [5]. This article combines this controller (will be referred to as BSB control) along with the FFT for excitation force estimation, and uses the resulting system to numerically simulate the amount of energy that can be converted using the SNL’s experimental buoy when subject to Bretschneider wave spectrums. The next section describes in some detail the FFT approach implemented in this article.

Fast Fourier Transform for wave estimation The fast Fourier transform method implements the Fourier transform concept to transform a series (sequence) of data points from the time domain to its representation in the frequency domain. In doing that, one can figure out the frequencies in the measured signal, the individual amplitude at each frequency, as well as the phase shift at each frequency.

In this study, we measure the buoy position and pressure at discrete points on the buoy surface. These measurements are collected over a time period (called a window); the data collected over this window are processed using the FFT methods to identify the signal in the frequency domain. The frequencies, amplitudes, and phases are then used to estimate the excitation force at the end of the time window. This excitation force is then used to compute the control force at the end of the

ASME Power Division: Plant Operations and Maintenance Committee

A message from the chair The Renewable Advanced Energy Systems (RAES) committee of ASME Power Division was created to address the growing interest with power generation from renewable energy sources papers and presentations. This Committee shall function to promote the art and science of power generation through professional publications, peers review, teleconferences and online discussions (https://www.linkedin. com/groups/7015325 ). This committee provides a forum for the benefit of those engineers, manufacturers, consultants and multidisciplinary researchers, who have indicated an interest in the areas relating to: power generation with renewable sources, advanced technology, combine heat and power or advanced steam/organic cycles.

The ocean waves and tidal are incessantly source of kinetic energy, more consistent than wind to integrate as base power in an electric grid. But offshore power conversion, whatsoever wave or wind source, is not inexpensive; the past European Pelamis’™ funding problem shows that perhaps wave’s energy flourishing years have yet to come.

In contrast, more of our American project design is based in reasonable cost WEC buoys as you can see in the wave energy prize final (http://waveenergyprize.org/teams) sponsored by the U.S. Department of Energy (DOE)‘s Water Power Program. In the ASME feature article in this issue, one of our active members presents numerical experiments on a buoy prototype. This study shows how to improve Wave Energy Conversion performance and efficiency by an improvement of through the control system design.

The Committee meets four times a year, with one of the meetings coincide with the ASME Power or the PowerGen International conferences. Our meetings, LinkedIn group and ASME web, are an opportunity to network with professionals with similar interests. The RAES is a open forum to exchange ideas and learn more. Our upcoming summer meeting will be held at the ASME Power & Energy Conference in Charlotte, North Carolina, June 26 to 30, 2016. We look forward to seeing you there!

Figure 1. Flow chart of the concept of combination of controller and FFT estimator

May 2016 | ASME Power Division Special Section

Reza Arghandeh Chair – ASME Power Division RAES Committee arghandehr@gmail.com

ENERGY-TECH.com

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ASME FEATURE window using a BSB control. So, a time window, moving in time, keeps updating the frequencies. A Hamming window is employed. The size of this window is a design parameter that can be tuned for best performance. The concept of the combination of FFT and the controller can be shown in the flow chart in Figure 1

Simulation results Numerical Simulations were conducted through numerically integrating the equations of motion, with radiation forces and excitation forces computed based on a finite element software that provides an accurate computation for these forces.

The measurements are simulated by adding noise to the simulated position and pressures at the sensors locations. The FFT algorithm uses these simulated measurements and estimates the excitation force. The BSB control algorithm then computes the necessary control to maximize the extracted energy. This section shows graphically the results of these numerical experiments. Two test cases are investigated. The first is a spherical buoy excited with a wave of only 3 frequencies. The second is a test case on the actual SNL’s experimental buoy, excited by a Bretschneider wave spectrum.

Figure 2. Position of the floater

Test case 1: Spherical buoy Consider the case of a spherical buoy of radius 1 m; the frequencies are selected to be 2π ,2π 2π rad/s __ __ , __ 2 3 5

The equilibrium level is such that half of the sphere is submerged. Hence, the mass of the sphere in this case is 2.0944×103 kg. The corresponding added mass at infinite frequency is 1.1253 × 103 kg. The Stiffness is 3.0819e+04N/m.

And the viscous damping coefficient is 17.4239N.s/m. Three radiation states are included in the system dynamics model.

Figure 2 shows the simulation results for the position of the buoy over time, and Figure 3 shows how the velocity changes over time. Each of the two figures has two lines; one of them is the true value as computed from the numerical simulation and the other one is the value as computed by the FFT algorithm.

The two values, in general, are different due to the noise in the measurements, errors in the models used, and inaccuracies in the FFT algorithm itself.

A more important metric of interest is the energy extracted. Figure 4 shows the extracted energy in this case. There are two traces. One trace shows the energy that would be extracted if there are no noise in the measurements or inaccuracies in the FFT algorithm. The second trace shows the simulated actual extracted energy when simulated errors are included. The two traces in the figure are close which indicates an acceptable performance for the proposed system.

There are several parameters that impact the extracted energy. The noise level in the measurements affects the extracted energy in the sense that these noises affect the accuracy of the estimated excitation force. To highlight the impact of noise

20 ENERGY-TECH.com

Figure 3. Velocity of the floater

Figure 4. The Energy extraxted by using BSB controller with FFT window is 35.5 s

level, the numerical experiments are repeated for several noise levels and the results are compared in Figure 5.

Another factor is the FFT window size. This parameter can also be tuned for better energy extraction. Figure 6 shows

ASME Power Division Special Section | May 2016

ASME FEATURE

Figure 5. The Energy comparison between different level of noise

Figure 8. The Control Force

Figure 6. The Energy comparison between different window sizes

Figure 9. The Energy comparison between Complex Conjugate control, BangSingular-Bang control and Resistive Loading control

Figure 7. The estimation for excitation force with the window size equal to 38 s

Figure 10. Wave elevation versus frequency

the extracted energy for few different values of window size to highlight the impact of this parameter.

To determine how efficient is the FFT process is in estimating the excitation force, Figure 7 shows the excitation force. Figure 8 displays the computed control force using BSB,

for both cases. The first case assumes perfect knowledge of the excitation force and the other case uses the estimated excitation force. Clearly, the two traces are close from each other which indicates that the errors in estimating the excitation force do not significantly impact the computed control.

May 2016 | ASME Power Division Special Section

ENERGY-TECH.com

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ASME FEATURE

Figure 11. Schematic of the SNL Experiment WEC with locations of pressure transducters

Figure 13. WEC system heaving point absorber composed of a floating buoy [7] design and fabricated at SNL

Figure 11, has a mass of 858:4 Kg, a volume of 0:8578 m3, and a diagonal inertia matrix of [83:9320;83:9320;137:5252] Kgm2.

It is assumed that there are eight pressure sensors on one quadrant of the buoy surface at different heights.

Another series of numerical results were generated that are similar to the previous section on the spherical buoy case. These are shown this time only for the extracted energy. Figure 12 shows the extracted energy assuming a window of 80 seconds.

Figure 12. The Energy extracted compared to the maximum theoretical limit for the SNL’s Buoy

Finally, to highlight the significance of the proposed control system it is compared to another system. The analysis of this problem in the frequency domain yields the complex conjugate control as a means to compute the maximum theoretical obtainable energy from a control system. Another base line control is the resistive loading control. In Figure 9, the proposed control system results are compared, with respect to extracted energy, to the theoretical limit as computed from the complex conjugate control as well as the resistive loading control, in terms of extracted energy. The energy extracted from the proposed BSB control is very close with that obtained from the complex conjugate control if there are no measurement errors.

When some errors are included, the extracted energy is slightly less but significantly higher than that of the resistive loading (RL).

Test case 2: Sandia’s experimental buoy The dynamic model in this case simulates the true model and assumes 32 frequencies in a bretschneider wave spectrum, as shown in Figure 10. The test cases considered assume the SNL experimental buoy configuration. The device, shown in 22 ENERGY-TECH.com

SNL’s experiment At the time of writing this article, an ongoing advanced controls project, conducted by SNL, to study the improvement of WEC performance through the control system design is being sponsored by the DOE’s Wind and Water Power Technologies office. Controllers were selected to span the WEC control design space with the aim of building a more

Figure 14. WEC system converts wave motion to electrical energy through a linear generator system (WEC in the NSWCCD MASK basin facility)

ASME Power Division Special Section | May 2016

ASME FEATURE comprehensive understanding of different controller capabilities and requirements. To design and evaluate these control strategies, a model scale test-bed WEC was designed for both numerical and experimental testing. A number of control strategies have been developed and applied on a numerical model of the selected WEC [6]. This model is capable of performing at a range of levels, spanning from a fully-linear realization to varying levels of nonlinearity. Experimental testing is a critical step in the development of models describing the behavior of a system. A WEC device test-bed was used to perform this study’s control performance comparison. In Figure 13, a photograph is shown of the device which was recently tested in the Naval Surface Warfare Center, Carderock Division (NSWCCD) Maneuvering and Seakeeping (MASK) basin. The objective of the experimental testing is to obtain models for the design of control systems [6] in realistic environments for a Wave Energy Converter (WEC). The particular WEC considered here is a heaving point absorber [7] composed of a floating buoy (see Figure 13) connected to a support structure through a linear actuator. The support structure is then attached to the side of a bridge (see Figure 14) [8]. The next steps in this ongoing project will be to evaluate the control strategies with the experimentally calibrated WEC model and compare with planned future MASK basin testing for next year.

Conclusions The numerical simulations conducted in this paper demonstrated that the singular arc optimal control of a single-degree of-freedom wave energy converter results in extracting energy levels that are comparable to the maximum theoretical limit predicted by the complex conjugate control; the match between the two levels of energy depends the measurements noise and the window size of the fast Fourier transform method. The extracted energy is significantly higher compared to that obtainable though the base line resistive loading control.

The fast Fourier transform demonstrated feasibility of estimating excitation forces for Bretschneider waves. ■ REFERENCES 1. [1] National oceanic and atmospheric administration, national data buoy center, http://www.ndbc.noaa.gov/hull.shtml.

2. [2] Fusco, F., and Ringwood, J., 2014. “Hierarchical robust controlof oscillating wave energy converters with uncertain dynamics”. Sustainable Energy, IEEE Transactions on, 5(3), July, pp. 958–966.

3. [3] Cummins,W., 1962. The impulse response function and ship motions. Schiffstechnik.

4. [4] Falnes, J., 2002. OceanWaves and Oscillating Systems - Linear Interactions Including Wave-Energy Extraction.

Cambridge University Press.

5. [5] Zou, S., Abdelkhalik, O., Robinett, R., Bacelli, G., and Wilson, D., submitted March 2016. “Optimal control of wave energy converters”. Renewable Energy, Elsevier, in review.

May 2016 | ASME Power Division Special Section

6. [6] Bacelli, G., Coe, R., Wilson, D., Abdelkhalik, O., Korde, U., Robinett, R., and Bull, D., 2016. “A comparison of wec control strategies for a linear wec model”.

In Proceedings of the 4th Marine Energy Technology Symposium METS2016.

7. [7] Bull, D., Coe, R., Monda, M., Dullea, K., Bacelli, G., and Patterson, D., 2015. “Design of a physical point-absorbing wec model on which multiple control strategies will be tested at large scale in the mask basin”. In International Offshore and Polar Engineering Conference (ISOPE2015).

8. [8] Wilson, D., Bacelli, G., Coe, R., Robinett, R., Thomas, G., Linehan, D., Newborn, D., and Quintero, M., 2016.

“Wec and support bridge control structural dynamic interaction analysis”. In Proceedings of the 4th Marine Energy Technology Symposium, METS2016.

Editor’s note: This paper, PWR2015-49032, was printed with permission from ASME and was edited from its original format.To purchase this paper in its original format or find more information, visit the ASME Digital Store at www.asme.org.

By Ossama Abdelkhalik, Shangyan Zou and Rush Robinett; Michigan Tech University, Department of Mechanical Engineering and Engineering Mechanics, Houghton, Michigan and Giorgio Bacelli and David Wilson; Sandia National Labs,Water Power Technologies and Electrical Sciences and Experiments Departments, Albuquerque, New Mexico.They can be reached at editorial@woodwardbizmedia.com.

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MACHINE DOCTOR

Gear tooth cavitation erosion

By Patrick J. Smith

The causes of gear tooth damage are not always obvious. Gear tooth damage can be costly and can result in excessive machine downtime. The purpose of this article is to present an unusual case study of gear tooth damage in an integrally geared turbocompressor.

note that the oil supply temperature was steady at 57째C during the event. After the trip, the lube oil increased to 61째C over a span of 3 minutes then cooled back off. The trip was attributed to an issue in the electrical substation and the machine was restarted and reloaded without incident.

Introduction This case study pertains to a dual service, 5 stage, integrally geared centrifugal compressor driven by a 3570 RPM, 3000 HP induction motor. The gearbox consists of a bullgear and three rotors. The low speed (LS) rotor operates 22,118 RPM and consists of a pinion with impellers mounted at each end.

The LS rotor comprises the first two stages of the main air compressor service. The high speed rotor (HS) operates at 31,875 RPM and also consists of a pinion with impellers mounted at each end. The HS rotor comprises the third stage of the main air compressor service and the first stage of the booster nitrogen compressor service. The cover rotor operates at 30,214 RPM and consists of a pinion with an impeller mounted at one end. The LS and HS rotors rotor are mounted on the gear case horizontal split line, while the cover rotor is installed in a split line in the upper gear case cover and shifted towards the HS rotor side of the gearbox. The compressor configuration is shown in Figure 1 and 2. As shown in Figure 2, oil is sprayed out of the gear mesh for each rotor.

One month later, the compressor tripped again following a restart after a short plant outage. The compressor was started without incident. Five minutes after the startup, the compressor was being loaded when the MAC 3rd stage vibration spiked up and out of range of the transmitter. A couple of minutes later the motor power spiked up again, similar to the previous incident and the motor tripped on high current. As before, the oil supply temperature increased after the trip before cooling back down.

The gearbox utilizes tilting pad journal bearings for both pinions and there is a single non-contacting proximity type vibration probe adjacent to each bearing except on the cover pinion where the probe measures on the OD of thrust collar.

The pinions are also fitted with thrust collars which are used to transmit pinion axial thrust to the bullgear. The thrust bearings are on the bullgear rotor as shown. The bullgear journal bearings are a cylindrical sleeve type and the thrust bearings are a tapered land type. There are no vibration probes on the bullgear rotor. The compressor protection system includes high pinion vibration alarms and high high pinion vibration shutdowns.

Two days later the compressor tripped on high MAC 3rd stage vibration. A review of the trends showed a small, slow increase in the oil supply temperature followed by a sharp spike in

History After the compressor was commissioned, it was put into continuous service. The machine was in operation for approximately 22 months when it shut down during normal, steady state operation on high motor current (power). A review of the trends showed that the motor power slowly increased by 5% over a 5 minute period even though the compressor was operating at constant process conditions. The motor power then suddenly spiked to 167% of the normal operating steady state power, tripping the motor on high current. It is interesting to

The compressor was inspected and there was significant damage to the bullgear thrust bearings, several pinion bearings, and there was LS pinion gear tooth damage on the tooth tips and unloaded flanks. Further review of the trends showed that the auxiliary oil pump (AOP) did not start when the compressor tripped due to a controls issue. It was thought that this contributed to the gearbox damage. The compressor was repaired, the control logic was updated and the compressor was restarted without incident.

Figure 1. Compressor Configuration

May 2016 ENERGY-TECH.com

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MACHINE DOCTOR The gearbox was rebuilt with new parts at the OEM shop.

The reassembly was closely followed to ensure there were no issues. This was followed by a comprehensive shop test and post test gear inspection. No problems were identified and the compressor was shipped and installed at site.

No problems were uncovered during the shop test. However, without any explanation for the odd gear wear or the other damage that seemed to coincide with the unusual mechanical behavior (sudden increase in motor load, vibration and post shutdown oil supply temperature increase), there was a concern that the problem was not solved.

Figure 2. Gear and Oil Spray Arrangement

motor power just before the vibration trip. As before, the oil temperature increased after the trip before cooling off.

The gearbox inspection covers were removed and there was no obvious gear tooth damage. There was a suspected problem with a motor contactor. This was fixed and the compressor was restarted again.

The compressor ran for another week before it tripped again on high motor power. A review of the trends showed a similar pattern. Prior to the vibration trip the oil supply temperature slowly increased followed by a sharp increase in motor power.

The trends are shown in Figure 3. An inspection through the gearbox inspection cover showed frosting like indications on all the gear teeth at the same end of each pinion. There was no visible damage on the corresponding bullgear teeth.

Based on the gear damage and the continued issues with power spikes, oil temperature excursions and high vibrations, the compressor gearbox was removed and sent to the OEM for detailed inspection. Other than the gear tooth wear, no other damage was found.

The pinion bearings, bullgear bearings and impellers were all in good condition.

Unfortunately the machine tripped after 48 hours of runtime in the same manner as before. There was a sudden rise in motor power which led to a shutdown, followed by an increase in the oil supply temperature after the trip. An inspection showed damage to the bullgear thrust bearings, some minor impeller rubs and some minor seal damage. There was also some minor tooth tip erosion on the gear tooth tips of all three pinions. The damage was not as bad as before, but the damage occurred in only 48 hours of runtime.

Corrective action All along it was suspected that there could be an issue with gearbox flooding whereby the oil doesn’t drain fast enough out of the gearbox, oil level builds up and then the bullgear starts pumping oil. This scenario would lead to an increase in power and would heat the oil, which over time would overwhelm the oil cooler resulting in warmer oil supply temperatures. This could also cause an additional thrust load which could explain the bullgear thrust bearing failures. In turn, this would open up the bullgear float which could then account for the impeller

A close visual examination of gear teeth showed that the frosting like indications were actually pit type defects and were present on the tooth tips at the disengaging end of every pinion gear tooth. The damage was about 0.5 inches from the end of the tooth and was on the tips and extended to a lesser degree down the unloaded flanks and into the roots. A picture of the damage is shown in Figure 4.

The compressor is a well referenced frame, although the two pinion configuration is more common than the three pinion configuration with the cover rotor. The oil spray arrangement and oil flows were also standard for this gearbox.

26 ENERGY-TECH.com

Figure 2. Trends May 2016

MACHINE DOCTOR rubs. To prevent future incidents and to further assist the investigation, the following changes were made: • Lowered the high motor current set point to trip the motor sooner if a problem develops • Added bullgear thrust bearing temperature probes with high temperature alarm and shutdown protection.

• Added a bullgear axial probe with high axial position alarm and shutdown protection.

• Added a gearbox vacuum pressure transmitter with low vacuum alarm and trip protection.

• Added temperature probes in the bottom of the gearbox to help determine oil is collecting in the sump during operation.

• Added a bubbler” level gauge on the gearbox sump as another means to help determine if oil is collecting in the sump during operation • Installed a larger vacuum blower to increase the gearbox vacuum to about 9 inches water column, which was consistent with other machines at this site.

• Replaced the drive motor with the spare. This was done just to eliminate any possible problem with the drive motor.

teeth including hardness, chemistry and microscopic exams.

No material defects were uncovered.

Applied load As previously described the damage to the gear teeth was on the top lands, unloaded flanks and in the roots. These are non-contacting regions and so by observation, it seemed very unlikely that applied load during normal operation was a factor in the gear tooth damage.

Installed a vacuum gauge on the gearbox cover for additional information.

Drive Systems Technology, Inc. was also contacted to perform a failure analysis on one of the failed pinions. It was decided to use the LS pinion.

Gear failure analysis The following causes were investigated as part of the LS pinion gear tooth failure analysis: • Material defect • Applied load • Corrosion • Stray electrical currents • Cavitation The LS pinion is comprised of 47 teeth and operates at a rotational speed of 21.142 RPM. The gear is made from AISI 4340E steel and the surface is nitride hardened to 47 – 55 Rc. Material defect Several dimensional checks were made followed by a magnetic particle examination.

No dimensional issues were found and no cracks were uncovered. The gear was then sectioned and numerous metallurgical tests were performed on pieces of the damaged

May 2016 ENERGY-TECH.com

27

MACHINE DOCTOR

Figure4. Gear Tooth Damage

Corrosion Close examination of the pits showed the areas to be clean with no evidence of corrosion or chemical attack.

Electrical discharge Electrical discharge damage is typically seen on the softer surfaces on Babbitt bearings, not on the harder gear tooth surfaces.

There were no signs of damage due to electrical discharge in the bearings. So, it seemed very unlikely that electrical discharge caused the gear tooth damage. In addition, part of the metallurgical exam included a nital etch inspection. This test did not show any evidence of arc burns, rehardening and/or retemper burns which would be consistent with electrical discharge. So, it seemed unlikely that electrical discharge caused the gear tooth damage.

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Cavitation A close visual inspection of the damaged areas showed pits that were consistent with cavitation damage. The damaged areas had a sponge like appearance and the material appeared to have been “eaten” away. This is consistent with the damage in pump impellers that have suffered cavitation damage. Cavitation is the formation of vapor bubbles in a liquid that occur when the local hydrostatic pressure drops below the vapor pressure of the fluid.

If this is followed by a rapid pressure increase, these bubbles implode and can generate an intense shock wave sufficient to damage metal surfaces.

Although there is no physical test that was done to prove the damage was due to cavitation, the lack of another cause and the appearance of the damage suggested that cavitation was most likely the cause.

Discussion Cavitation damage in gear pumps has been known to occur, but cavitation erosion in high speed, helical gears is virtually unheard of. It was speculated that this could be due to an issue in the lubricant supply to the gear mesh or possibly a consequence of gearbox flooding. Since the last repair, the compressor has operated without incident for over three years. The only change was that the gearbox vacuum was increased. One possibility is that the additional vacuum helps better drain the oil by pulling it into the reservoir. Due to the past problems, there was no incentive to try testing at different vacuum levels to see if the problems return due to concerns that this could cause machine damage.

There is an identical compressor and another similar compressor at the same site. After the problems with the compressor discussed above, the high amp shutdown set points were lowered on these other machines. In addition, the bullgear thrust bearing temperature protection, bullgear axial position protection and the gearbox vacuum pressure protection were retrofitted. A couple of years after the problems with the above compressor, it was noted that these other compressors had problems with trips due to high motor amps whenever both the main oil pump (MOP) and auxiliary oil pump (AOP) May 2016

were operated at the same time. When both of these pumps are operated, there is a higher oil pressure and hence higher oil flow to the gearbox. It is possible that this aggravates a problem with oil drainage from the gearbox. Fortunately, there has been no gear tooth damage and it is very likely that the lower motor amp setting tripped the machine before damage occurred.

REFERENCES 1. 1. Drago, Raymond J., Cunningham, Roy J., and Flynn, William, “The Anatomy of a Lubrication Erosion Failure: Causation, Initiation, Progression and Prevention, Part 1,�

Gear Solutions, March 2014 2. 2. Drago, Raymond J., Cunningham, Roy J., and Flynn, William, “The Anatomy of a Lubrication Erosion Failure: Causation, Initiation, Progression and Prevention, Part 2,�

Gear Solutions, April 2014

To address this additional gearbox oil drains were added in the side of the gearbox of one of these machines near the bottom and in the side inspection cover adjacent to the LS pinion. This was done to address any oil collecting in the bottom of the Patrick J Smith is lead machinery engineer at Air Products & Chemicals in Allentown, Pa., where he provides technical machinery gearbox as well as any oil that collects in the cover pinion area support to the1 company’s air separation, hydrogen 1511-SCHENCK_1-2i.pdf 10/13/15 operating 10:33 AM due to crowding in the gearbox under the cover pinion and processing and cogeneration plants. You may contact him by emailing windage that prevents proper drainage. Since making these editorial@woodwardbizmedia.com. changes, there have been no issues with operating the compressor with both oil pumps in operation.

Conclusions The failure described in this article was very unusual. It is very likely that the machine failures were due to problems with oil drainage. Gear spray arrangements, oil flow requirements, and drainage are largely based on supplier experience. What was challenging in this case was that this was a well referenced frame. It is likely that some subtle differences with other similar machines were enough of a change to cause a problem with gearbox oil flooding. The testing with the additional drain lines confirms a flooding issue.

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The gear tooth cavitation erosion damage is even more unusual and it is also very likely that this damage was a consequence of gearbox oil flooding. The inspection and testing done by Drive Systems Technology, Inc was very useful for the failure analysis and these types of inspections should be considered whenever there is unusual gear tooth damage.

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Author’s Note: A colleague will be presenting a case study on the oil drain modifications at the 2016 Turbomachinery Symposium in Houston, TX. Interested readers are encouraged to attend the symposium to learn more about how these modifications were designed, implemented and tested. â–

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