Solving Thrust Bearing Overheating Problems.

For several years, Grant County Public Utility District (PUD) struggled with high temperatures – and subsequent failures – of the thrust bearings of two units at its 907-MW Priest Rapids hydro project. To solve the problem, the utility installed pressure transducers to measure load on the bearings, then used the measurements to more accurately adjust the bearings’ position.

Project background

The Priest Rapids project is on the mid-Columbia River in Washington. The powerhouse is equipped with ten vertical Kaplan turbines. Each umbrella-style generator is rated at 95 MW. Unit commissioning started in 1959 and was completed in 1961.

English Electric designed and manufactured the thrust bearing assembly for each unit. This assembly consists of ten tilting pad shoes supported with an equalizing table. Each shoe is pre-loaded with an adjusting screw. The thrust runner is in two pieces bolted together, and then bolted to a thrust collar. The thrust collar is shrunk fit 0.05-inch onto the generator shaft.

During start up and commissioning, it was necessary to scrape a depression in the babbitt to prevent overheating of the bearings. The scrape patterns evolved through the years with minor bearing wipes, developing their own patterns of areas needing a depression scraped. Often, a bearing set would require a wear-in period and more than one scraping before the temperatures settled. As a result of frequent high temperatures in the thrust bearing assembly, Grant County PUD conducted an inspection and scraping every four years.

Modifying the thrust runner and bearing

In 1995, after several bearing failures, PUD engineers decided to investigate alternatives to hand scraping for solving the overheating problem. On two units they measured bearing movement relative to the support, pressure at the high lift port, and temperature distribution across the leading and trailing edges of the thrust bearing shoe. From an analysis of the test results, the engineers concluded the oil wedge between the thrust runner and thrust bearing (also referred to as an oil film) pressurized the split between the two thrust runner halves, causing it to open slightly. This opening – aided by centrifugal force – allowed oil to flow out the end of the split. The oil leak resulted in unequal load on the thrust bearing, allowing the outer radius of the bearing to move up and briefly make contact with the thrust runner. This contact caused the temperature to rise on the outer radius and across the top of the thrust bearing shoe. As a result of the temperature increase, the bearing deformed into a crowned shape; this caused the top of the bearing to wipe if a depression was not scraped into it.

PUD engineers conducted testing of the thrust bearings and thrust runner to resolve the overheating and failure issues. During data collection, a physical observation of the thrust runner split leak confirmed this unusual phenomenon. PUD crews reported that, when the unit was rotated manually (with the high-pressure lift system energized), a stream of oil “squirted” out the split in the thrust runner. Fretting corrosion between the thrust runner and thrust collar near the runner split was attributed to the split becoming pressurized and moving the thrust runner slightly.

As a result of the analyses, the thrust runner on every unit was removed and the thrust runner split machined to achieve a tight fit. The connecting bolts on each thrust runner were shortened to provide more resistance to flexing. All the thrust bearing sets were machined flat without the scraped depression. Another modification involved replacement of a threaded plug for the high-pressure lift system port with a plug that was welded flush. Additionally, all the radial anchor grooves – which were intended to help hold the babbitt in place but can be a source of bonding problems – were machined flat.

From 1995 to 2004, as a result of the modifications to the thrust runner and bearing, PUD was able to discontinue the previously described four-year cycle of scraping, inspection, and maintenance on the thrust bearings. The number of forced outages caused by thrust bearing problems improved from an average of almost one unplanned outage a year to one unplanned outage every three years.

Recent failures

Then, in December 2004, a bearing failure occurred on Unit 1. Three of the ten bearing shoes had sections of babbitt completely removed. This failure was attributed to babbit bond failure on an older rebabbitting process. In the areas of babbitt removal, the bearing shoe had no tin left. The babbitt was previously bonded to the shoe with a trimetal copper process and anchor grooves.

In May 2005, the Unit 1 thrust bearing failed again. A visual inspection showed the babbitt contained fatigue cracking on all the bearings shoes. PUD staff discovered three of the eight bolts between the thrust runner and thrust collar had broken. They also found a crack in the thrust runner split joint, resulting from the additional stress caused by the broken bolts. Powertech Labs performed material failure analysis, consisting of micrographs and scanning electron micrograph pictures. This analysis showed that the three bolts failed in tension due to fatigue.

After this second failure, PUD staff disassembled the entire bearing support mechanism in Unit 1, including the equalizing table, to inspect it for wear or failures. On the surfaces of the thrust runner and thrust collar where they face each other, staff found excessive fretting corrosion. The corrosion caused the thrust collar to be out of tolerance.

To fix the problem, PUD staff would need to machine the thrust collar in place. However, staff concluded in-place machining would be too risky. It would require design and fabrication of a custom machine tool. The high original tolerances for flatness on the thrust collar would be difficult for a custom machine tool. If the flatness tolerance was degraded further during an in-place machine operation, complete disassembly would be required to repair the damage.

The next option – to dismantle the unit even further to machine the thrust collar using a standard available machine tool – was too costly. This option would result in lost power generation from the unit for about two months.

The PUD ultimately decided to replace the thrust bearing runner with a spare and put the unit back in service. The PUD decided the damaged and out-of-tolerance thrust collar would not be repaired at this time.

Before placing the unit back on line, the PUD added monitoring instrumentation to provide feedback on the operating characteristics. PUD personnel placed a resistance temperature detector (RTD) inside the six thrust bearings with no monitoring instrumentation. The PUD already used these RTDs on four of the bearings to measure temperature in the thrust bearing shoe.

Figure 1: Grant County Public Utility District connected pressure transducers to each of ten bearings in Unit 1 at its Priest Rapids hydro project to measure

oil pressure at the high-pressure lift port.

In addition, they connected a pressure transducer to each bearing to measure the oil pressure at the high-pressure lift port, as shown in Figure 1 on page 72. The end nut of the bearing was drilled and tapped to allow the pressure at the port of the high-pressure lift system to be measured while the unit was on line. This modification does carry the risk of developing leaks between the oil port and the pressure transducer. A significant leak in the tubing could lead to the loss of the oil wedge and result in a bearing wipe. The risk was minimized by careful installation and pressure testing of all the tubing.

Grant County PUD had one month to assemble the unit and put it into operation, to meet future power demand. The short time available prohibited the use of a less risky, more traditional method of measuring the pressure, such as load cells or submersible transducers installed close to (but not on) the bearing.

In September 2005, the PUD began placing Unit 1 back in service. First, the thrust bearings were pre-loaded with the adjustment screw that holds the thrust bearing shoe up against the thrust runner, according to standard torquing procedures. However, two of the bearings did not leak oil out of the edges of the bearing with the high-pressure lift system energized. If a bearing shoe does not leak oil, the load on that particular bearing is too high. In addition, the unit would not rotate manually.

A second attempt was made at torquing the adjusting screws, yet the results remained the same. The PUD staff had no success with additional attempts to flood the bearing by jogging the pump, raising the pressure relief on the pump, and manually

rotating the unit with winches. They were able to get the unit to rotate, albeit with difficulty, by lowering two of the thrust bearings by 0.001-inch. Although the problems were not solved, the unit was placed back in service because the outage schedule had been exhausted.

When the unit was in operation, the pressure fluctuation on every pad (measured by the transducers) looked similar to a sinusoidal wave function with a frequency of twice per revolution. During manual rotation while the unit was off line, PUD staff determined the pressure was lowest at the point when the thrust runner split was located directly over the high-pressure lift port (the pressure transducer port). Conversely, the pressure was highest when the split was 90 degrees from the port.

Staff recorded the minimum, maximum, and average value of each pressure transducer. The difference between the minimum and maximum value represented the level of pressure fluctuation on each pad during one revolution. The average value for each pressure transducer was used to qualitatively compare the load on each of the ten thrust bearing pads relative to each other.

When Unit 1 was running, the average difference in bearing pressure was 1,770 pounds per square inch (psi), with the highest average bearing pressure around 3,000 psi and the lowest around 1,230 psi. The average difference between the minimum and maximum values on a single pad was 400 psi.

After 18 days of operation, high thrust bearing temperatures developed, and the unit had to be taken out of service. PUD staff lowered the two highest bearings and raised the two lowest bearings until the bearing pressures (with the high pressure system operating) were closer in magnitude.

After the second adjustment, the average difference between bearing recordings was 514 psi with the high-pressure lift system running, and 1,111 psi when Unit 1 was on line. The 1,111 psi was an improvement from the previous value of 1,770 psi. Average operating temperature of the bearings was reduced by 2 degrees Celsius (C) under the same approximate operating head and generator load.

This tubing on a turbine-generating unit at Grant County Public Utility District’s Priest Rapids project connects to the

high-pressure lift port through the

thrust bearing end nut and, ultimately, to the pressure transducer.

In September 2006, Unit 7’s thrust bearings wiped. During the bearing replacement, pressure transducers were installed on each thrust bearing so that pressure measurements could be compared to Unit 1. The thrust bearings in Unit 7 were pre-loaded with the standard torquing procedure and then adjusted up and down using the following process, developed by PUD personnel:

First, the high-pressure lift system was energized and the unit was rotated until all the bearing pressures stabilized. Then, the average pressure for each bearing during one revolution was calculated. Third, the unit was raised off the thrust bearings, and several designated bearings were lowered or raised based on the average values. The highest and lowest pressure thrust bearings were adjusted first. Fourth, the high-pressure lift system was de-energized, thus lowering the unit back down on the bearings. Then, the pressure in each thrust bearing was recorded again. This process was repeated until the difference in average pressure could no longer be improved.

This process does take longer (a full day vs. the previous half day of preloading only). However, it provides the ability to ensure the load is equalized across all ten thrust bearings. The more equally distributed the total load is on the bearing, the better the operating characteristics. However, this also highlights a drawback to the pressure transducers vs. load cells. With a load cell, Grant County PUD most likely could adjust the bearings in one step. The pressure transducer value cannot be recorded while making the adjustment, so the entire process was necessary in order to adjust the bearings.

The pressure data collected from Unit 7 was compared to that of Unit 1. The average difference in bearing pressure on Unit 7 was 300 psi while running compared to the 1,111 psi on Unit 1. The average bearing pressures (with the high lift system operating) for Unit 7 differed by 400 psi between the high and low bearing when they first pre-loaded and were improved to an average bearing pressure of 80 psi between the high and low bearing. Operating under the same approximate head and load, the highest bearing pressure on Unit 7 was 1,600 psi while on Unit 1 it was 2,500 psi. The data recorded from Unit 7 confirmed that the thrust bearing adjustment could still be improved on Unit 1.

In December 2006, the PUD conducted a planned outage of Unit 1 to make a third adjustment on the thrust bearings, using the same process as it used on Unit 7. The average difference in bearing pressures was decreased to 182 psi compared to 514 psi with the high-pressure lift system operating only, and 742 psi compared to 1,111 psi with the unit in operation. The adjustment resulted in an average operating temperatures reduction of 0.5 degrees C by evening out the load on the bearings.

Although the evenness of the load distribution in Unit 1 improved significantly from the first adjustment (from 1,770 psi to 742 psi), the improvement did not allow the unit to match the average bearing pressures in Unit 7. Further observation of the data revealed that the average bearing pressures from Unit 1 have a wider spread when the unit is loaded with hydraulic thrust than those on Unit 7. Some of the Unit 1 bearings do

not pick up an equal share of the load. The reason for this discrepancy has not been determined.

Units 1 and 7 are now running satisfactorily. During an upcoming planned outage, PUD staff will close a valve immediately downstream of the drilled port in all ten bearing assemblies in Unit 1 to further minimize the risk of leaking oil. The units are scheduled for a major rehabilitation in five to ten years.

Lessons learned

Good written record keeping is important for finding solutions to recurring problems. For the Priest Rapids thrust bearing problems, finding historical information took significant effort, and some information had been lost. During a failure involving an unplanned unit outage (especially a reoccurring failure), the focus is often on returning the unit to service and not determining the cause of failure. Yet, observations noted during a failure often help with future problems. Grant County PUD has improved its record-keeping process by archiving the information found so far using Maximo from MRO Software and producing internal reports on recent failures.

The pressure transducers have turned out to be a cost-effective alternative to load cells for measuring the bearing load. They cost less, are quicker to obtain, and are easier to install. Originally, Grant County PUD elected to use the transducers rather than load cells owing to a limited time for instrumentation. To date, the transducers have provided accurate data, useful in making bearing adjustments.

The addition of more temperature data collection and oil pressure data on Unit 1 led to a new adjustment procedure for the troubled unit, which has allowed it to operate without continued thrust bearing failures. The PUD believes the new procedure for thrust bearing adjustment is superior to the old process of torquing that will result in more equal bearing loading and longer bearing life.

Teflon coated thrust bearing

The babbitt lining can be replaced by teflon lining. During 1995, at the time of up gradation of Bhakra Right Bank Power Plant machines, M/s TPE Russia, supplied Teflon coated thrust bearing pads. This reduced thrust bearing temperature by about 10 C. These pads are being used in some Western Countries since 1974. Its other advantages are –

i). The brake application speed can be further reduced, thus decreasing wear of brake pads & hence contamination of winding.

ii). The pads are less sensitive to scratches.iii). Due to less friction, the need for high oil pressure at the time of starting &

stopping may be eliminated.

Thrust bearings support axial loads on rotating shafts. Designs range from simple, coin-sized flat washers in household appliances to sophisticated assemblies several feet in diameter for hydroelectric generators.

Six basic types are available. The first, an externally pressurized, hydrostatic thrust bearing, works for low-speed, heavily loaded equipment including telescopes,

observatory domes, and large radio antennas, where structures may weigh a million pounds or more.

Hydrostatic thrust bearings use an external pump to provide oil-film pressure when simple, internal hydrodynamic pumping action cannot generate sufficient force. Primary use is in equipment run at extremely low speeds, under high loads, with low viscosity fluids, or where space is limited. A compact thrust bearing can feed high-pressure oil into a single pocket at the end of a rotor, for example. Larger bearings may employ three or more pressurized pockets. Hydraulic flow resistors in the supply line to each pocket, or equal flow to each pocket from ganged gear pumps, provide the asymmetric pocket pressures needed to support off-center loads. Unit loading on such bearings is usually limited to about 0.5 to 0.75 × external pump feed pressure of up to about 5,000 psi.

The other five thrust bearing types internally generate oil pressure (self-acting) to support thrust loads. Here, a rotating face or shaft collar pumps oil onto a supporting thrust-bearing surface.

Tapered-land thrust bearings find use in mid to large-sized high-speed machines such as turbines, compressors, and pumps. In most designs, a flat land extends an additional 10 to 20% of the circumferential breadth B at the trailing edge of each segment. This extension can boost load capacity 10 to 15% and reduce wear during starts, stops, and at low speeds. Gradual wear increases this flat portion to about 30 to 50% of total area, which helps maintain load capacity. In many turbine and compressor applications, individual segments are square (radial length L = B) and have a circumferential taper of about 0.003B0.5.

Tapered-land bearings are sensitive to load, speed, and lubricant viscosity, and therefore are commonly designed to match operating conditions of specific, constant-speed machines.

Pivoted-pad thrust bearings are typically used in turbines, compressors, pumps, as well as marine drives, in much the same general size and load range as tapered-land designs. Pads automatically adjust to form a nearly optimal oil wedge that supports high loads over widely varying speeds in either direction and with a variety of lubricants. Leveling links behind the pivots accommodate minor misalignment and equalize loads on each of three to 10 pads. Most units contain six pads, with outside diameters twice the inside diameters. Slot-shaped oil inlet openings between individual pads consume about 15% of available area between the inside and outside diameters.

Offsetting the pivot location about 65% beyond the leading edge raises load capacity, lowers operating temperatures, and cuts power loss. Replacing steel with copper for backing of the babbitt bearing material also lowers peak surface temperature. Oil fed directly into a leading edge groove in each pad (nonflooded lubrication) minimizes hot oil carryover from pad to pad. It also lets oil drain from the housing to mostly eliminate parasitic power loss at high surface speeds. Pivot location is usually set 55 to 58% radially outward on the pad to avoid radial tilt.

Film thickness is minimal with low-viscosity fluids such as water, liquid metals, and gases. In such applications, pads incorporate a small spherical or cylindrical crown with a height 0.5 to 2× the minimum film thickness. The arrangement handles loads about

equal to flat-surfaced pads that have an optimum pivot location. The downside: Bearings with offset pivots rotate in one direction only.

Spring-mounted thrust bearings are some of the largest self-acting types, carrying millions of pounds in hydroelectric generators, for example. Each pad mounts on a nest of precompressed springs to avoid the high contact stresses otherwise imposed by loading individual pivots. In smaller bearings where axial space is at a premium, rubber backing provides the flexible support.

Spring-mounted bearings typically run at speeds from 50 to 700 rpm at projected unit loads of 400 to 500 psi. While individual pads are often square (L/B = 1), the largest diameter bearings use elongated pads with B shorter than L. The shortened path in the tangential direction of motion avoids overheating the oil film and babbitt bearing surface.

These large spring-mounted bearings are built to tight tolerance, which helps maintain a thin oil film during starts and stops, and provides ample oil-film thickness for continuous operation.

Step thrust bearings use a coined or etched step. As such, they are well suited to mass-produced small bearings and thrust washers. They work with low-viscosity fluids such as water, gasoline, and solvents. Step height must nearly equal minimum film thickness for optimum load capacity, yet be large enough to permit some wear. A step provides the same amount of hydrodynamic pumping action as a wedge, though the stepped design hasn't caught on for large machinery because it tends to accumulate dirt. Wear and erosion diminish step effectiveness.

Flat-land thrust bearings are the simplest and least expensive to make. They handle light loads for simple positioning of rotors in electric motors, appliances, crankshafts, and other machinery. Flat-land bearings carry 10 to 20% the load of other thrust-bearing types. This is because flat parallel surfaces do not directly build oil-film pressure through pumping action. They depend instead on thermal expansion of both the oil film and bearing surface to generate an oil-supporting wedge.

Small flat-land bearings with no oil-distributing grooves handle unit loads from 20 to 35 psi. In larger bearings, adding four to eight radial oil-distributing grooves improves oil feed and cooling, raising unit load to about 100 psi.

MATERIALSTin babbitt (typically ASTM B23, Alloy 2: 88% tin, 7.5% antimony, 3.5% copper)

gets the nod for most industrial, marine, and transportation equipment. The material resists corrosion and helps prevent scoring of rotating steel thrust surfaces because hard dirt and wear particles easily embed into its surface. Applying a thin tin-babbitt layer — a few mils thick on a bronze or steel shell, up to about 125 mils thick on larger units — partially offsets the material's low fatigue strength with oscillating loads. Applying a thin electroplated babbitt overlay to a copper alloy substrate helps avoid transfer of the latter to steel thrust runners.

Lead babbitt (typically ASTM B23, Alloy 15: 83% lead, 15% antimony, 1% arsenic, 1% tin) costs less than tin babbitt. Use well-inhibited lubricating oil to avoid corrosion by oxidized oil, especially with water contamination.

Leaded bronzes (83% copper, 7% tin, 7% lead, 3% zinc) are in many small and low-speed machines as low-cost thrust washers and bushing thrust faces.

Reinforced plastics and porous iron and bronze work for bearings and thrust washers in fractional horsepower motors, appliances, and automobile and agriculture

equipment. Carbon graphite and rubber work for bearings run in water and various low-viscosity fluids.

Upgrading Thrust Bearings at Akosombo

Installation of new runners at the 1,020 MW Akosombo project revealed the inability of the existing thrust bearings to handle the higher hydraulic thrust. Replacing these bearings with a new runner plate and polytetrafluoroethylene (PTFE) bearings significantly reduced unit operating temperatures and eliminated recurring cavitation damage.

In the mid-1990s, the Akosombo Generating Station, on the Volta River in Ghana, contained six turbine-generating units with a total capacity of 912 MW. The first four units, commissioned in 1965, had a capacity of 588 MW. The final two units were commissioned in 1972 and had a capacity of 324 MW.

After 30 years of operation, owners the Volta River Authority carried out a technical audit with the objective of modernizing the plant to operate for another 25 to 30 years. This audit indicated that the units could also be upgraded to increase power production at the facility by nearly 12%.

In 1999, Volta River Authority began a major overhaul of the generators and turbines that included installation of new runners supplied by GE of Canada and VA Tech of Austria. This upgrade increased station efficiency and capacity. Now, Akosombo has a capacity of 1,020 MW.

However, installing the new runners at the Akosombo plant caused some problems. The new design increased hydraulic thrust while operating at the same flows as the previous runner, contributing to rare recurring cavitation damage on the generator thrust bearings.

After the upgrade, thrust bearing operating temperatures increased to 86oC from 76oC. To rectify the situation, the original equipment manufacturer (OEM) recommended rearranging the thrust bearing support springs. This reduced the high temperatures to an average of 82oC, but a residual problem of recurring cavitation damage to the thrust bearing pads emerged.

After several unsuccessful attempts to remedy the problem, Volta River Authority contracted Hydro Tech Inc. of Canada to design a new thrust bearing featuring a modified highly-toleranced split rotating ring, keyed and bolted together to form one complete ring (with no gaps at the splits). This would solve the cavitation problem, but the bearing would still be overloaded. Thus, Hydro Tech decided to improve load capacity of the bearing pads by replacing them with polytetrafluoroethylene (PTFE) pads.

Installation of the new thrust bearing assembly was completed on Units 2 and 5 at Akosombo in November 2009. Operating temperature was reduced 7oC to 8oC (from 83oC to 76oC). Continual monitoring of thrust bearing temperatures indicates good operation of the units.

Discovering the thrust bearing problem

Before the new runners were installed at Akosombo, Volta River Authority and its consultant, Canada-based Hatch, raised concerns about the possibility of increased thrust load and the ability of the existing bearings to match this load. However, the OEM assured them that the thrust bearing was capable of supporting the new levels of thrust loading.

Unit 3, the first to be overhauled at Akosombo, operated with an average thrust bearing temperature of 76oC before the upgrade. During commissioning of the upgraded unit in July 2000, this temperature increased to an average of 86oC, confirming increased thrust load associated with installation of the new runner. Monitoring was carried out while the retrofit project was under way. The Unit 3 thrust bearings failed while the unit was operating in April 2002, after almost 20 months of service. Even though other circumstances were involved, such as irregular cooling water flows, this failure again raised concerns about the existing thrust bearings.

Volta River Authority contacted the bearing OEM to study the problem and offer an appropriate solution. The OEM carried out a bearing optimization study at its own cost and recommended a revised support spring arrangement to increase the oil film thickness, therefore providing more reliable bearing operation with reduced temperature. The OEM implemented this solution by removing 11 support springs, mainly from the leading edge end of the thrust bearing. This resulted in a total of 40 support springs as compared to the original 51 springs per bearing pad. After this modification, the average thrust bearing operating temperature for Unit 3 dropped to 82oC.

The new polytetrafluoroethylene (PTFE) bearing pads being installed on the six units at the 1,020 MW Akosombo Generating Station have reduced unit operating temperatures to acceptable levels.

Although thrust bearing operating temperatures were now reduced, a new problem of rather rapidly recurring cavitation damage became obvious on the thrust pad surfaces. This cavitation damage was first discovered during an inspection six weeks after installation. The damage was mostly concentrated within and around the high-pressure oil injection groove ring, toward the trailing end of the bearing pad.

Cavitation of thrust bearing pads is unusual and typically is caused by a design flaw. At Akosombo, the highest measured average bearing pressure is about 4.07 MPa, but there are higher and lower pressures on the bearing surface. Generally, the middle of the bearing pad surface toward the trailing edge, close to the high-pressure oil lift port, is a higher load area. This port is where the most significant pressure drop occurs because of the change in oil film thickness and/or pressure. This results from a step in the babbitted surface and a gap between the thrust bearing runner plate segments.

With the increased thrust, the oil film between the babbitt thrust bearing pads and runner plate decreases, while average pressure rises. This increases the pressure drop at the thrust bearing runner plate split and increases its effect on the babbitt due to close proximity, resulting in cavitation pitting of the pads.

During a thrust bearing inspection performed in August 2008, evidence of fatigue failure of the babbitt surfaces was discovered, likely due to thermal effects (a condition known as “thermal ratcheting”). Thermal ratcheting is caused by temperature changes on the bearing surface, which causes the babbitt bearing pad to act as a bi-metal strip, slightly bending the bearing pad and causing crowning. Crowning occurs because the babbitt expands and contracts at a different coefficient of expansion to steel, and the surface of the bearing pad is hotter than the steel backing plate to which the babbitt is bonded. (The thicker the babbitt, the stronger the thermal ratcheting effect.)

When bolted together, the new runner plate on this Akosombo unit features no gaps or splits, thus eliminating pressure disturbances that can cause cavitation damage to the bearings.

These unusual bearing pad damages became a greater concern to Volta River Authority and its contractors as cavitation and thermal ratcheting progressed steadily. The OEM experimented with several modifications by machining different tapers and shapes within the high-pressure oil lift groove ring, but the problem persisted. It became obvious that initial concerns about the ability of the bearings to safely support increased thrust load associated with the new replacement runners were valid.

Developing a solution

Volta River Authority, working with Hatch, met with the OEM for a permanent solution to the Akosombo thrust bearing problem. The OEM’s preferred solution was to install a new one-piece rotating ring, an approach that involved completely removing the generator rotor. But because of the prolonged outage time involved with this approach (at least three months) and the fact that the thrust load might still be above the safe load-carrying capacity of the existing babbitt thrust pads, Volta River Authority did not accept this solution.

Volta River Authority contracted Hydro Tech to modify the Akosombo generator thrust bearing assembly to deal with the recurring cavitation problem. All parties fundamentally attributed this problem to the increased hydraulic thrust load of the new turbine runner. However, the original bearing was already susceptible to cavitation due to gaps in the runner plate quadrants. Before the turbine upgrade, there was at least one incident of cavitation damage discovered on a bearing pad. At that time, it was considered to be a result of porosity in the babbitt.

Objectives of the Akosombo generator thrust bearin g retrofit were to:

- Address the fact that the thrust load of the new runners was above the safe load limit of the existing thrust bearings; - Solve the design flaw that was causing the cavitation on the thrust pad surfaces; and, - Lower operating temperatures.

Increasing thrust load

Hydro Tech first measured thrust loads on the units to inform the design of the new PTFE thrust pads. Thrust was as high as 4.07 MPa at a headwater elevation of 247 ft (75 meters) above sea level (maximum headwater level at Akosombo is 278 ft (285 meters) above sea level). Actual maximum thrust load increases as the headwater level rises. Full head pond thrust measurements could not be taken because the pond does not completely fill each year. The results again indicated that total thrust on the units was above the safe operating limit of babbitt bearings.

PTFE bearing pads with higher load capacity were installed to replace the traditional babbitt bearings. The new PTFE thrust pads for Akosombo feature compound tapers at the leading edge to facilitate oil flow and wedge formation, as well as pad eccentricity to permit higher specific load capacity. Using the PTFE bearing pads also allowed Hydro Tech to eliminate the high-pressure oil lift system, removing another potential source for cavitation.

Solving the design flaw

The solution Hydro Tech provided to address sudden pressure fluctuations caused by gaps at the splits of the existing rotating rings was a modified design of highly-toleranced segmented rotating rings, which are keyed and bolted together to ensure no gaps. This type of joining was achieved during installation. The splits were lapped (stoned) to completely eliminate sharp ends, eventually forming a one-piece thrust bearing rotating ring that was bolted to the thrust block.

The new thrust bearing runner plate acts as a solid one-piece ring, providing constant pressure on the bearing pads over its entire surface.

Lowering the operating temperature

The PTFE bearing has one-fifth the coefficient of friction compared with the existing system, preventing excess heating of the thrust oil. This allows the thrust bearing temperature to be lowered with no additional cooling.

Results

With installation of the new thrust bearings on Units 2 and 5, all problems that resulted in higher operating temperatures have been resolved. The bearing pads were inspected after several months of running time and appeared to be in good condition. Bearings on the remaining units will be changed as soon as an outage time is available.

Sticky Wickets: Seal Ring Installation Device Shortens Outage Time during an Overhaul

During the recent overhaul of its 663-MW Shasta plant, the U.S. Department of the Interior's Bureau of Reclamation needed a method to quickly and efficiently replace the stationary seal rings. Personnel developed a device that would allow the seal rings to be in the turbine pit with the headcover and boring bar installed. The seal rings then were installed and machined while the contractor worked on the stator rewind overhead. This device was used on the final three of the five units overhauled at this plant and reduced total outage time by several hundred man-hours.

Developing the device

Reclamation owns 83 hydroelectric plants. For the most part, Reclamation uses its own staff to assemble and disassemble the 131 turbine-generator units in its plants. The use of Reclamation staff has resulted in some remarkable innovations. In particular, one such innovation at the Shasta plant, on the Upper Sacramento River in California, significantly reduced the outage time during a recent unit overhaul.

The five units at Shasta feature vertical shaft generators. Original plant maintenance practices relied almost exclusively on the plant main gantry crane to remove the generator and turbine components during disassembly and reassembly. Traditionally, these units were disassembled from the top down, and parts were staged throughout the plant.

In 2007, Reclamation staff finished the complete replacement of the second Shasta turbine assembly. Alstom, the contractor Reclamation selected for the turbine replacement, determined that the turbine assembly efficiency could be improved by replacing the turbine, wicket gates, and operating linkages. The contract was awarded

for all five main units at Shasta. For two of the five units at Shasta, this overhaul also included a full stator rewind. Normal practice was to completely tear down the unit, from the pilot exciter down to the draft tube. Then the lower guide bearing bracket was reinstalled, and a work platform was erected above it for the generator rewind. Once the rewind was completed, the work platform and lower guide bearing bracket would be removed.

For those units that had already been rewound by Alstom, the practice would be to remove the headcover, invert it, then install the upper seal ring on the power plant deck. The lower seal ring would be installed separately in the discharge ring. The headcover and boring bar would be installed and the two rings turned to the final dimension.

The Shasta turbine replacement involved installation of new stationary seal rings. This was needed because the original seal rings, designed in a segmented style, had a history of failures and seizures. Seal rings, also called wear rings, reduce the water that either bypasses the turbine blades or applies force to the top of the turbine, thus increasing the thrust bearing load. Alstom redesigned the seal rings for the Shasta units into a continuous style. This new design would require that the seal rings be installed with an interference fit to the headcover and discharge ring. Figure 1 shows the design and location of typical seal rings, much like the new Shasta seal rings.

Personnel at the 663-MW Shasta plant prepare to use a specially-designed device to install the upper seal ring in one of the turbine pits. Replacing the seal rings in three units with the headcover and boring bar installed reduced total outage time by several hundred man-hours.

Once plant personnel removed the work platform for the rewind, the new seal rings normally would have to be lowered into the turbine pit and installed. Once installed, the seal rings would need to be machined, which requires the installation of a boring bar that would rely on the headcover for its upper support.

In preparation for the turbine replacement for Unit 4, performed in 2002, a plant mechanic named John Boughton had the idea that the seal rings could be in the turbine pit with the headcover and boring bar installed. In this situation, the two seal rings could

be installed and machined in one set without having to either turn the headcover upside down or remove the headcover. The idea turned out to be very beneficial when Alstom personnel worked on the stator for Units 1 and 2. Boughton sketched out his idea. Owing to his other duties, Boughton was not able to fully develop the design.

Another plant mechanic, John Martin, moved ahead with the concept and designed the device. The device needed to be designed so that it could be assembled under the headcover and would support the seal rings in preparation for installation. The device had to fit around the boring bar, which was centered in order to machine the seal ring to the final dimension. The device resembled a twin-arm bearing bracket, with two long elements roughly equal to the diameter of the smallest seal ring. The device featured six bolted segments. Two of these segments separated the long elements, and four more (perpendicular to the two long elements) extended the line of the two separating elements to form a symmetrical cross. All the elements were made of 2-by-6-inch channel and featured bolted joints.

This design contained a 3-inch gap for dry ice to shrink the diameter of the seal ring. With this design, the seal rings were supported by six bolts that would be threaded through a gusset welded to the end of a 13-inch-long channel that was 2 inches wide by 4 inches deep. This arrangement also would support an internal micrometer to measure the change in diameter of the seal ring.

During fabrication of the device in 2002, Martin altered the design, assisted by John Hays, to replace the 13-inch-long channel with a flat plate and threaded block. This plate would be used to support the seal ring. A bolt through the block was reduced in size from the original design and now used only to center the ring. Figure 2 shows the final design of the device inside a cross-section of a discharge ring and the lower seal ring.

The device Martin designed could accommodate two different upper and lower seal ring diameters and ensured the seal rings could be centered and measured for the proper fit.

One particular benefit is that the smaller-diameter seal ring could be installed with little modification of the device. The design relies on the symmetry of the bracket to rotate the support plate end for end, move the plate to the opposite arm, and reverse the jack bolt. With that change, the other-diameter seal ring could be installed. Unfortunately, owing the different radius, another set of dry ice trays had to be fabricated.

The entire assembly could be removed through the scroll case access hatches once both seal rings were installed.

Using the new device

This device was used on all Shasta units, including the final two unit rewinds and turbine replacements. Reclamation personnel estimate that the device reduced the outage time by several hundred man-hours compared with installing the new seal rings after the rewind was completed. This equates to about $40,000 in savings due to reduced labor costs alone.

This device can be adapted to any unit that is undergoing seal ring replacement during a rewind. The device is particularly suited for power plants with larger turbine diameters and similarly-designed units.

One final note: The device was improved on several times during its evolution. John Martin took the lead in the evolution of the device, but the ideas and suggestions provided by others that resulted in further improvements were not recorded. The entire Shasta maintenance and engineering staff undoubtedly contributed and should be recognized. Finally, credit should be given to the management personnel in the Shasta office, who continue to encourage this sort of drive and innovation.

Repairing a Failed Turbine: The Story of Returning G.M. Shrum Unit 3 to Service

When a major failure occurred in Unit 3 at G.M. Shrum, one of the largest hydro plants in Canada, BC Hydro's objective was to get the machine back on line as quickly as possible. Through the use of innovative repair methods, the unit was rebuilt and back on line in 14 months. The repair at this 2,730-MW plant is an example of the innovation and technical expertise abounding in the North American hydro market ... no matter the unit size.

On March 2, 2008, the Unit 3 runner at the 2,730-MW G.M. Shrum Generating Station experienced a major failure. This failure resulted in significant damage to the runner and water passage components, and operation of the unit ceased until repair work could be performed.

G.M. Shrum, BC Hydro's largest facility, began operating in 1968 with ten turbine-generating units. Units 1 to 5 were installed in 1968 and 1969 and have a capacity of 261 MW each. These Francis turbine runners are fully cast from mild carbon steel and have a stainless steel overlay for cavitation protection. Together, these five units represent 12 percent of BC Hydro's electricity-producing capacity. Units 6 to 8 were installed in 1971 and upgraded in 2004 and have a capacity of 275 MW each. Units 9 and 10 were installed in 1974 and 1980, respectively, and have a capacity of 300 MW each.

Because of the importance of Unit 3 at G.M. Shrum, BC Hydro implemented a plan to get it back on line as quickly as possible. This plan involved rebuilding the runner by using a new blade design and reusing the existing crown and band, as well as implementing a novel approach to repair the damaged wicket gates. Using these innovations, BC Hydro was able to return Unit 3 to service in just 14 months.

How and why the unit failed

The failure event began at 5:05 a.m. on March 2, 2008, when unit output dropped to 237 MW from 254 MW. The servomotors began moving open, to 84.3 percent from 71.3 percent. Power output returned to 254 MW at 5:06:27 a.m. By about 5:11 a.m., the servomotors had reached 97.2 percent open, with power output of 262 MW. Turbine bearing temperature and synchronous vibration began increasing.

Then, at about 5:14 a.m., power output dropped to 247 MW before settling at 252 MW. Turbine synchronous vibration increased again. At 5:19:30 a.m., power output increased to 258 MW. Between about 5:21 and 5:22 a.m., the servomotors moved to 93 percent open from 97.2 percent open, with no change in power output. At about 5:22 a.m., power output plummeted to 13 MW, without servomotor movement. Output then settled at 55 MW, and turbine inlet pressure stabilized at a value 3.7 percent higher than nominal. Turbine bearing temperature began climbing rapidly.

At 5:28:30 a.m., the operator tripped the unit and power output dropped to 0 MW. One minute later, the unit reached 133.9 percent speed because the wickets gates were no longer able to regulate the flow of water through the turbine. Two minutes after that, turbine bearing temperature reached 129.4 degrees Celsius (C). Finally, at about 5:39 a.m., the unit stopped.

In the days after the failure event, BC Hydro personnel performed a unit assessment. Below are results of this assessment, as well as subsequent inspections throughout the disassembly process:

– The scroll case and stay vanes were free of damage from impact or abrasion. There was no indication of material or debris having traveled down the scroll case or passed through the stay vanes, and no residual debris was found in the scroll case.

– The skin plate on the pressure side of all 24 wicket gates was in good condition. However, the skin plate on the suction side, downstream of the seal contact line, had deep gouges on up to 75 percent of the surface. The pattern of the damage was indicative of pieces of metal becoming caught in the wicket gate/runner cascade while the turbine was rotating.

– The trailing edges of wicket gates 1 to 10 and 15 to 24 were bent due to impact. On all 20 of these gates, the impact pattern and location of the damage were the same (about 12 inches above the bottom facing plate). However, the trailing edges of wicket gates 11 to 14 were in good condition, with no impact damage.

– Runner blades 4, 11, and 14 were missing significant pieces (about 70 inches by 70 inches) from the outlet. The shapes of these missing pieces were almost identical, and one piece was found intact in the draft tube. In addition, many smaller runner blade pieces (8 inches by 8 inches) were found in the runner/wicket gate cascade, stuck between adjacent blades, and in the draft tube.

– The inlet edge of all 17 runner blades had impact and heavy abrasion damage. On "moderately" damaged inlet edges, the damage was localized at 12 inches above the bottom facing plate (the same elevation as the trailing edge damage on wicket gates 1 to 10 and 15 to 24). In addition, significant cracks were visible on 12 of the blades.

The Unit 3 runner at 2,730-MW G.M. Shrum suffered significant damage during a failure event, including a large piece missing (see arrow) from the inlet edge of one of the runner blades

– The area between wicket gates 11 to 14 contained a carbon steel deposit on the stainless steel lower seal ring, indicating that rubbing had taken place with the runner band.

– All shear pins had failed, with pin 11 failing due to fatigue, pin 13 failing due to bending at a high rate of strain, and all other pins failing due to shear overload.

The repaired Unit 3 runner for 2,730-MW G.M. Shrum was delivered to the facility in February 2009, just 11 months after the failure.

BC Hydro concluded that the most likely failure mode was a cascade closure of four adjacent wicket gates due to shear pin failure. In this scenario, the shear pin on wicket gate 11 failed due to fatigue, resulting in rapid closing of the wicket gate. Shear pin 11 functions as a shear pin and link pin. Thus, this element would experience cyclical bending and torsional forces due to increased clearances between the pin and bushing, lubrication issues, misalignments in the operating mechanism, and interference between the lever and mechanical stop on the head cover.

The lever of wicket gate 11 then contacted the lever of wicket gate 12, breaking its shear pin. Both wicket gates assumed an almost closed position. The governor then instructed the servomotors to increase the wicket gate opening to maintain power output and to respond to an increased demand for power. Shortly after the servomotor reached its maximum opening, the shear pin on gate 13 failed. Wicket gate 13 closed and its lever contacted that of gate 14, breaking its shear pin.

With four wicket gates closed, water flow in the scroll case became unbalanced, and a low-pressure zone was created behind those gates. The runner was forced toward the low pressure until contact occurred between the runner band and lower seal ring. The stresses on the runner caused new cracks and accelerated existing cracks, resulting in failure of blades 4, 11, and 14. As the blades broke apart, one piece was discharged into the draft tube and two were ejected into the wicket gate/runner cascade, slamming the remaining wicket gates closed, breaking the remaining shear pins, and further damaging the runner blades and wicket gates.

Choosing the repair method

On March 13, 2008, BC Hydro invited three turbine suppliers to view the damaged turbine. On March 20, BC Hydro issued a request for proposals to these suppliers. In its proposal, each supplier was to describe its repair method for each turbine component, including the cost and expected duration. Proposals were received on March 31. After reviewing the technical, schedule, and cost aspects of each proposal, BC Hydro determined that the solution offered by Voith Hydro was the best for its business needs.

On April 15, 2008, BC Hydro issued a notice to proceed to Voith Hydro for repair of the Unit 3 turbine. This turbine is scheduled to be upgraded in 2017 as part of the G.M. Shrum Units 1 to 5 Turbine Upgrade Project. Thus, the goal of the repair work was to

return Unit 3 to commercial operation as quickly as possible, with a minimum service life of ten years.

The runner repair was the most time-intensive work needed to get the turbine operational. To expedite the repair, Voith Hydro proposed to reuse the crown and band and supply new blades. This removed the long lead time for crown and band castings, which would have added about one year to the schedule. To further reduce the time required for the repair, Voith Hydro chose to fabricate the new blades using A516 Gr. 70 plate steel instead of stainless steel. This plate steel was a stock item readily available from a warehouse, whereas the stainless steel would require a special delivery. Thus, the repair cycle time was reduced to nine months.

However, the decision to reuse the crown and band with a new blade design presented challenges. For example, the new blades were designed using only finite element analysis (FEA) and computational fluid dynamics (CFD). No model testing was performed. Additionally, the new blade design had to fit within the existing crown and band water passageway, including leaving locations for the thrust relief holes and space to install new rotating wearing rings on the outside diameter.

For the wicket gate repair, the main goal was to retrieve the functionality of the gates. A time frame of seven months was foreseen for this work.

The wicket gates are of the hollow fabricated type, made of welded carbon steel plates. The body is made of two formed 1.75-inch-thick plates with an internal rib welded on the suction and pressure side plates. The nose contact face is equipped with a rubber strip to obtain a tight seal when closed, held in place with a stainless steel clamp plate and patch bolts. The head cover and bottom ring also feature wicket gate end seals so the distributor is watertight when closed. Additionally, the top and bottom end of each gate body is overlaid with 17-7 stainless steel.

Once the unit was dismantled, BC Hydro personnel discovered that the damage on the wicket gates was extensive. The suction side of the hydraulic surface showed scores in the range of 0.625 inch deep over its entire surface. Some wicket gates also showed deformations of the suction side skin plate, with cracks through the thickness of the plate. Cavitation damage was present immediately downstream of the nose rubber seal and in the vicinity of the stem to blade fillets. Under normal conditions, the wicket gates would have been replaced. However, due to time constraints, this was not an option.

BC Hydro and Voith Hydro worked together to identify three repair options:

– Completely replace the skin plate on the suction side;

– Weld overlay and grind the skin plate to restore shape and surface finish; or

– Mill a 0.5-inch-deep "pocket" on the suction side and install a new 0.5-inch-thick shaped skin plate. Damage deeper than 0.5 inch would be weld repaired before installing the skin plate.

The two first options were not desirable because they would induce significant distortions due to the welding volume involved. Therefore, BC Hydro and Voith Hydro determined that the third option was the most viable choice.

Making the repairs

Voith Hydro began disassembling the unit at the end of April 2008, only two weeks after the contract award. Disassembly activities were completed by the end of May 2008.

Runner repair

Immediately after the runner was removed from the turbine pit, Voith Hydro personnel removed the blades, leaving only 3 inches of blade protruding from the base metal. Voith Hydro then shipped the crown and band to its facility in York, Pa., arriving on July 8, 2008.

Because of the large amount of stainless steel overlay on the crown and band, Voith Hydro machined a 0.25-inch-deep cut on the crown and band water passage surfaces. The new blade design took this variation into account. The design also incorporated an overlay on the flange face and spigot area to keep the runner centerline at the distributor centerline (relative to the shaft flange face). After completing rough machining on the water passage surfaces, Voith Hydro personnel performed a magnetic particle inspection. This inspection confirmed that some of the cracking from the damaged blades extended into the crown.

The crown required extensive weld repairs. There were four defects exceeding 3 inches in depth and 18 inches long that were repaired with E309L weld metal. Machining on the crown revealed large areas of cavitation damage and stainless steel overlay, so Voith Hydro personnel machined away a layer of metal 0.625 inch deep in the area of the junction of the crown and blade discharge area. Voith Hydro personnel then overlaid this area with E309L filler metal. This area was then subjected to a non-destructive ultrasonic inspection. In total, the crown required about 1,800 pounds of weld metal. This included overlay and crack and cavitation repairs. The crown required re-machining after the weld repairs and before going to runner assembly.

The band repairs were not as extensive. The band only required about 100 pounds of weld metal and had no major defects. Only cavitation damage and the stainless steel overlay areas needed to be repaired.

Voith Hydro then completed runner assembly. The blade-to-crown and blade-to-band welds were made with E309L weld wire. These welds were inspected using non-destructive dye penetrant and ultrasonic methods. The completed runner was not post-weld heat treated due to the E309L welding. With the previous design, both the crown and band rotating seals had been integral. With the new runner, material was machined off of these areas to install new shrunk-on rotating seal rings. This allowed the seals to run true to the shaft spigot after the fabrication processes. Finally, the coupling holes were reamed oversized by 0.06 inch using a template to match the shaft, and new coupling bolts were installed.

The new runner has only 15 blades, compared to 17 for the old runner. Even with the decrease in number of blades, the total area of contact between the blades and crown increased by about 20 percent.

To be confident in the reliability of the repaired runner, Voith Hydro performed a complete FEA. This included a CFD study, static stress analysis, and dynamic stress

analysis. The dynamic analysis was performed to determine the risk of structural resonance in the runner and to evaluate its dynamic response over the range of exciting frequencies experienced during normal operation. The results from these studies were used for a fatigue study. Results from the fatigue study indicated that the runner will be reliable for more than ten years.

The replacement runner arrived at the site on February 20, 2009.

Wicket gate repair

In early June 2008, the wicket gates were shipped from G.M. Shrum to the shop of a Voith Hydro subcontractor for a complete rehabilitation. Shop personnel began with a complete examination of the wicket gates to map any cracks and abnormal conditions. Many cracks were repaired on the skin plates and also on the stainless steel overlay at both ends of the wicket gate body. Afterwards, a 0.5-inch-deep pocket was milled in the 1.75-inch-thick skin plate on the suction side where the gouges and dents were present. Subcontractor personnel then installed a formed and machined carbon steel plate in the pocket, restoring a clean and uniform hydraulic shape. The carbon steel plate was secured by means of a peripheral weld in addition to 16 patch bolts. The wicket gates went through a post-weld heat treatment before machining work began.

This wicket gate from Unit 3 at the 2,730-MW G.M Shrum plant shows typical damage uncovered after the runner failure. The skin plate on the suction side of all 24 wicket gates had deep gouges on up to 75 percent of the metal surface.

The remaining rehab work consisted of replacing the stainless steel sleeves with oversized ones, which provided the machining allowance needed to correct stem runouts. The contact faces of the wicket gate body had been re-machined to restore their parallelism. Therefore, it was necessary to re-dowel the gate arm on the wicket gate stem. The rubber seal and its clamp plate also were replaced with new ones. Finally, two coats of epoxy paint were applied on the hydraulic surfaces.

The wicket gates at 2,730-MW G.M. Shrum were repaired by milling a 0.5-inch-deep pocket in the 1.75-inch-thick skin plate. The contractor then welded and screwed a formed and machined carbon steel plate into the pocket to fill the cavity.

The structural modification of the wicket gates was validated using an extensive FEA model. The weld in the periphery of the new 0.5-inch skin plate was sized to restore the inertia of the gate body, while the size and number of patch bolts provided for an adequate clamping of the skin plate when subjected to vacuum (worst cases considered).

The wicket gates were returned to G.M. Shrum in January 2009, after only seven months of repair work.

Repairing other structures

The other structures repaired on Unit 3 include the embedded parts, head cover, turbine guide bearing, and turbine shaft.

Cracks on multiple stay vanes were repaired, and the embedded parts were machined at G.M. Shrum to restore an acceptable flatness.

Examination of the head cover showed only a few cracks that were quickly repaired. However, the geometry of the head cover needed reworking, which was accomplished by a Voith Hydro subcontractor.

The turbine guide bearing was heavily damaged. Rehab work consisted of melting the original babbitt, pouring a new coat of babbitt, and completely re-machining the coupling flanges and inside and outside diameters.

The turbine shaft was shipped to the shop of a Voith Hydro subcontractor for machining of the bearing journal and shaft seal journal surfaces.

Results

On February 20, 2009, Voith Hydro began reassembling the unit. Seven weeks later, the turbine-generator was ready to be commissioned by BC Hydro. Unit 3 was returned to service on May 5, 2009, just 14 months after the failure. To date, the repaired turbine has performed reliably and has met BC Hydro's expectations. In August 2009, minor cavitation damage was observed on the inlet edge of the runner blades. Voith Hydro implemented stainless steel overlay to repair the damage and to mitigate the effect of future cavitation.

As a result of lessons learned from the Unit 3 failure, BC Hydro has reviewed the risk of reoccurrence of this event on all turbines within its generating fleet. Included with this review is a reevaluation of BC Hydro's requirements for shear pin breakage monitoring and unit automatic shut down after a broken shear pin.