The purpose of this article is to present a case study of a challenging balance machine problem. A steam turbine rotor was sent to a machine shop for a standard low-speed balance during a planned turbine overhaul. Prior to the overhaul the turbine vibration levels were higher than desired, but had been stable for 6 years. This shop was very experienced with rotor balancing, and a turbine OEM representative was present to witness the balance job. But the problems encountered were not anticipated and the troubleshooting efforts led to the determination that the root cause was an issue with the setup of the balance machines being used. The purpose of this article is to describe the problem, the difficulties encountered during balancing and the solution.
This case study pertains to an industrial steam turbine that is direct coupled to a 51 MW generator.
The turbine is comprised of 19 stages and operates at 3,000 rpm. The turbine is fitted with drive end (DE) and non-drive end (NDE) “X” and “Y” shaft radial
The first major overhaul was performed in January 2005. During the overhaul, significant first- and second-stage blade rubs were discovered. The blade shrouds rubbed hard against the radial seals in the diaphragms, resulting in a significant amount of wear. This was a surprise because there was not any unstable vibration activity ever reported during operation. A picture of the damage is shown in Figure 1. In addition, two small dings were discovered on stage two 19 blades. This was presumably caused by foreign object damage.
It is highly likely that this condition existed for some time. During the overhaul, the rubbed/damaged surfaces were blended out and only a few grams of material were removed. The rotor was shot grit blasted and subsequently magnetic particle examined. No crack indications or other damage was discovered. The rotor was not dynamically balanced and was then re-installed in the turbine.
Following the overhaul, there were significant problems with high vibrations during ramp up when the turbine was accelerated through the 1st critical speed. There were also high steady state vibrations once the turbine was at running speed. Several balance weights were added to address the higher vibrations at both conditions. Virtually all of the vibrations were synchronous with rotor operating speed. In the end, a compromise was made to allow higher vibration levels during ramp up in order to achieve lower vibrations at full running speed. The maximum vibrations with the turbine accelerating through the 1st critical speed were 12.8 mils p-p (DE) and 8.2 mils p-p (NDE).
The maximum vibrations at running speed were 3.8 mils p-p and 3.1 mils p-p. This is described in the article entitled, “Significant Dates in Steam Turbine Vibration – A Case Study,” in the October 2009 issue of Energy-Tech magazine.
As described in the previous article, although the ramp up and steady stage vibrations were higher after the 2005 overhaul, it was decided to allow the machine to operate since the vibration levels at running speed were stable and in compliance with ISO standard 7919. However, the turbine controls were set to ramp on the fastest possible setting to accelerate the turbine through the critical speed as quickly as possible. Following the overhaul, the vibration behavior did not change significantly during transient startups and also was stable during steady state operation leading up to the next planned overhaul in July/August 2011.
After the rotor was removed from the casing, it was shot grit blasted and subsequently magnetic particle examined with no significant indications or findings. The first- and second-stage shroud damage that was observed in 2005 had not deteriorated further so the buckets were not replaced.