ENHANCED ROTORDYNAMICS FOR HIGH POWER CRYOGENIC TURBINE GENERATORS

The 9th International Symposium on Transport Phenomena and Dynamics of Rotating Machinery Honolulu, Hawaii, February -1, ENHANCED ROTORDYNAMICS FOR HIGH POWER CRYOGENIC TURBINE GENERATORS Joel V. Madison Ebara International Corporation 35 Salomon Circle Sparks, NV 93 Telephone: (775) 35-79 Fax: (775) 35- E-mail: jmadison@ebaraintl.com ABSTRACT This paper presents novel generator and thrust balancing configurations to enhance the rotordynamic behavior of Vertical Cryogenic Turbine Generators. The LNG production plants currently in the design stages are utilizing economy of scale to increase production with lower capital costs, which has the effect of increasing the power ratings of the critical rotating equipment being supplied. The higher power ratings of the variable speed hydraulic turbine generators implicitly require larger physical sizes, however the possible loss of generator load demands critical speed separation below minimum continuous speed and above maximum "breakaway" speed. The limitations on bearing design, shaft size and overall diameter of machines of this type impose certain restrictions on the design, which entail unique fabrication techniques to produce a robust, reliable machine. INTRODUCTION Figure 1 shows a conventional vertical cryogenic turbine generator that is rated at 115 kw power output. It can be seen that the turbine, consisting of three hydraulic stages, and generator share a common shaft. The turbine generator is mounted in a stainless steel pressure containment vessel, suspended by a flanged connection with a flexible gasket at the headplate. The LNG flows in from the top and expands through three radial inflow runners, which are mounted on a common shaft with the submerged generator. A small percentage of the high pressure inlet flow is routed through the thrust equalizing mechanism to balance the axial thrust load, then fed through the generator air gap to dissipate any heat production. The shaft is supported by three ball bearings, one each at the upper and lower ends of the generator and one at the turbine end of the shaft. Each bearing is lubricated by a controlled flow of process fluid, and is supported by the turbine casings. Circumferentially grooved seals are located between each stage and at the front shroud of each runner. The fluid routed through the generator air gap behaves as a hydrodynamic bearing due to the low pressure difference in the axial direction. The rotordynamic performance of this machine has been proven through analysis and performance testing data in the past (Habets and Kimmel; Smalley, Hollingsworth, Habets and Kimmel). However, as the power requirements increase with the demands for the next generation of liquefaction plants, the generator size must increase physically. As the generator is the dominant component in terms of rotordynamic stability, the increased size has negative ramifications to the machine reliability. As described in detail in previous work (Madison), the possible loss of generator load demands critical speed separation below minimum continuous speed and above maximum "breakaway" speed. For the turbine hydraulic combinations used in these applications the normal breakaway speed is on the order of 3 rpm. The calculated critical speed of the machine should have a Figure 1: Vertical Cryogenic Turbine Generator 1

nominal 15% margin above the maximum expected speed to ensure that no assumptions regarding the speed or the calculations will result in an inadequate design. This yields a minimum acceptable critical speed of 95 rpm. There are several unique aspects related to the design of this machine that are required as a result of operating while submerged in a cryogenic fluid (Weisser and Madison). These include limitations on bearing size, shaft diameter and generator diameter, all of which make it difficult to increase the physical size of the generator without detracting from the rotordynamic integrity. This paper presents solutions for increasing the generator size in small and large increments while still maintaining acceptable rotordynamic characteristics. PROPOSED INCREMENTAL ENHANCEMENTS For generator power increases of up to % simply reducing the span between the main bearings can offset the reduction in the critical speed resulting from the increased generator size. Referring again to Figure 1, it can be noted that this is not readily accomplished. The upper bearing needs to be located outside of the generator end turns to avoid any possible electrical effects associated from the variable speed electrical control system of the generator, and the lower bearing is integral to the thrust balancing system. One distinctive feature of these machines is that the deep groove ball bearings are cooled and lubricated by the product fluid. Due to the low viscosity of liquefied hydrocarbon gases the thrust generated by the hydraulic components must be totally eliminated for the bearing to achieve adequate running time between overhauls. While a device employing a combination of fixed and variable orifices has been utilized in the past, it has limitations with regards to location, length and resultant variable orifice gap size. variable orifice gap to increase to such as degree that the device would be rendered inoperable. The solution to this problem is to integrate another component within the device composed of a material that shrinks less than the shaft. The individual heights of the thrust plate and compensator portion are selected to produce the desired variable orifice gap at the actual operating temperature that will result in the most effective thrust balancing and highest machine efficiency. The rotordynamic performance of the original configuration and the modified machine were calculated using the methodology described in references 3, 5 and and compared to the original results. The analytical model geometry for the original configuration is shown in figure 3. The runner and generator masses are depicted by the attached weights, and the hydrodynamic and ball bearing supports are depicted by the spring symbols. Details on the hydrodynamic bearing stiffness and damping calculations can be found in the previous references. The ball bearings are located at 1.3, 37.9 and 9.5 inches from the left end of the shaft in the original model, yielding a bearing span of 5. inches for the generator and 3. inches for the turbine. The damped resonance frequency of the generator for this configuration was calculated to be 579 rpm, as indicated in the unbalance response plot taken at the center of the generator rotor, Figure. The deflected shape at the resonant frequency shown in Figure 5 confirms that the generator rotor first mode is causing the resonance. The revised geometry shown in Figure was employed for the second analysis case. As shown in Figure, the bearings for this geometry are located at 1.3,. and 9.5 inches respectively. This results in a reduction in the critical bearing span of.5 inches, or 5%. The span between the bottom guide bearing and the lower generator bearing is increased from 3. to 39.1 inches, however this segment of the rotor assembly has a second order effect on the critical speed as it has small masses that are well supported. As expected from the unbalance response results shown in Figure 7, the resultant critical speed was increased by % from 553 rpm to 55 rpm. The deflected shape at the resonant speed is shown in Figure, where it can be easily noted again that the generator is the source of the resonance. Increases in generator power cause a nearly linear increase in the physical length, which will result in a corresponding linear reduction in the critical speed. The revised configuration for the thrust balancing geometry can effectively compensate for an increase in generator size of up to approximately %, which will be acceptable for a large proportion of new projects. Figure : Comparison of Original and Modified Thrust Balancing Configurations Relocating the lower bearing by the requisite amount would entail a dramatic increase in the length of the thrust balancing device, as shown in Figure. The resultant differential shrink between the thrust plate material and the shaft would cause the NON-INCREMENTAL ENHANCEMENTS For increases in capacity beyond the % incremental limit demonstrated above, more novel techniques must be employed. One primary factor that drives a more concrete change in the configuration is that the generator frame size will necessarily increase in addition to the length. When the frame size increases, the rotor diameter will also increase, and centrifugal forces on the rotor itself become a concern in addition to the impact on the

35 Rotordynamic Response Plot Sta. No. : ROTOR CENTER 3 5 15 5 Excitation = 1x 1 Figure : Unbalance Response for Original Configuration < > Shaft Radius, inches 3 - - -3 Axial Location, inches Figure 5: Deflected Shape for Original Configuration.11..5.3. -.3 -.5 -. -.11 Rotordynamic Deflected Shape Plot, mils pk 3

5 35 3 5 15 5 Rotordynamic Response Plot Sta. No. : ROTOR1 CENTER 1 Figure 7: Unbalance Response for Revised Configuration 1, Modified Thrust Balancing Device 3.9 Shaft Radius, inches - - -3 Axial Location, inches Figure : Deflected Shape for Revised Configuration 1, Modified Thrust Balancing Device.7.5.. -. -.5 -.7 -.9 Rotordynamic Deflected Shape Plot, mils pk

critical speed. Fortunately, the variable speed controller allows for multiple generators to be controlled from the same source. This allows the single large submerged generator to be replaced with two smaller generators on a common shaft that are driven by the same set of hydraulics. The generators can be configured at either end of the hydraulic components or at the same end. The analytical model for the case with two generators located on the upper end of the machine with the turbine at the lower end is depicted in Figure 9. Each of the generators in this configuration is rated at exactly half the output power of the original single generator used in the baseline case. The actual length of each generator stack is 5% of the original single generator length for this duty. Figure 9 demonstrates that the three ball bearing system employed in configurations 1 and is replaced by a four bearing system, a guide bearing at the bottom of the shaft, a main bearing between the turbine and the first generator, a main bearing between the two generators and a main bearing at the upper end of the shaft. The revised configuration 3 now consists of three important bearing spans; 3.5 inches for the turbine hydraulics, 7.9 for the first generator and 35. for the second generator. While the overall length of the system has increased from 9.5 to 1.9 inches, the rotating masses have been more evenly distributed over shorter bearing spans, which will result in a more stable system. The unbalance response for configuration 3 taken at the center of the first rotor, shown in Figure 9, indicates that the critical speed has been raised to 75 rpm. This dramatic increase is to be expected considering the distribution of the masses and the shorter critical bearing spans. The resonance at 75 rpm is clearly attributed to the first rotor in this case, as a second resonance is also noted at 975 rpm, which is due to the second generator rotor, as depicted in Figure. Figure 11 shows the deflected shape of the rotor assembly at the first resonance, which indicates that the mode is being driven by the first generator, or middle mass element. Again, this result is to be expected considering that this element has the largest bearing span with the highest mass. The third alternate configuration being considered places a generator on either side of the turbine, as shown in Figure 1. As in the previous configuration the combination of the two generators gives the identical power rating to the original single generator configuration. This revised configuration also uses a four bearing system, but in this case all four bearings can be considered as main generator support bearings with the turbine connected between the two generators. The bearing spans for this configuration are 37 inches for the first generator, 31. inches for the turbine hydraulics and 3. inches for the second generator. With this configuration the end turns of the two generators were not located adjacent to each other so the overall shaft length only increased to 1.9 inches. The unbalance response results for configuration are shown in Figure 11, which indicate the critical speed to be approximately rpm. This increase is to be expected due to the reduced bearing span over the first generator as compared to configuration 3. The deflected mode shape for this case shows that both generators appear to be resonating at nearly the same frequency, which is an expected result given the similar bearing span over the two components. Again, as expected this configuration offers the most stability due to the more compact bearing spans and the resultant higher critical speed. SUMMARY The proposed changes to the configuration of the turbine generators show that increased capacity can be realized without sacrificing the rotordynamic integrity of the machine. Table 1 summarizes the resultant critical speeds and separation margins with respect to the expected maximum breakaway speed. Considering that there is a 9% separation margin built in to the baseline configuration, capacity increases of % over the baseline configuration can be accommodated through modifications to the thrust balancing system. This implies that the original design will be stable for power requirements of up to 13 kw. An additional benefit that can be realized with the revised thrust balancing system is the ability to tune the variable orifice gap to maximize the efficiency and stability of the machine. Configuration Critical Speed % Increase Over Baseline Separation Margin Baseline 553 rpm N/A 9% Revised Thrust 55 rpm % 3% Balancing System Two Generators 75 rpm % 3% Located Adjacently Two generators Located on Opposite Ends rpm 59% % Table 1. Comparison of Calculated Critical Speeds and Separation Margins for Each of the Four Analyzed Configurations For power outputs beyond % a two generator system is proposed. This could reasonably double the power output of the machine to ratings of 3 kw or higher while maintaining rotordynamic and mechanical integrity. The control and operation of the machine would not be complicated as the variable speed controller is fully capable of operating two generators in series. The modification to the trhust balancing system described in configuration is applicable to the two generator system also. An additional benefit to this configuration would be the ability to shut one generator down while operating at part load conditions, thereby operating the remaining generator closer to the rated point where peak efficiency and stability are attained. Of the systems proposed the preferred solution for higher powered units for future application is the configuration employing a generator on either side of the turbine. This will provide the most stability and will result in the highest reliability over the lifetime of the machine. 5

Rotordynamic Response Plot 1 Sta. No. : ROTOR1 Center 1 1 1 Figure 9: Unbalance Response for Revised Configuration, Two Generators on Same Side of Turbine, Center of Rotor 1 1 Rotordynamic Response Plot Sta. No.5 : ROTOR1 Center 1 1 1 Figure : Unbalance Response for Revised Configuration, Two Generators on Same Side of Turbine, Center of Rotor

Shaft Radius, inches 3 - - -3-1 Axial Location, inches Figure 11: Deflected Shape for Revised Configuration, Two Generators on Same Side of Turbine..... - - - - Rotordynamic Deflected Shape Mils pk 1 Rotordynam ic Res ponse Plot Sta. No.11 : ROTOR1 CENTER 1 1 Figure 13: Unbalance Response for Revised Configuration 3, Two Generators on Opposite Sides of Turbine, Center of Rotor 1 7

1 Rotordynamic Response Plot Sta. No.5 : ROTOR CENTER REFERENCES Habets, G.L.G.M., and Kimmel, H. E., April 1999, 1 Figure 13: Unbalance Response for Revised Configuration 3, Two Generators on Opposite Sides of Turbine, Center of Rotor Shaft Radius, inches 3 - - -3-1 Axial Location, inches.3..1.1. -.1 -.1 -. -.3 Figure 1: Deflected Shape for Revised Configuration 3, Two Generators on Opposite Sides of Turbine Development of a Hydraulic Turbine in Liquefied Natural Gas, Proceedings of the 7th International Fluid Machinery Conference, IMechE, The Hague, The Netherlands. Rotordynamic Deflected Shape Plot, mils pk Weisser, G. L. and Madison, J.V., Application of Aerospace research on Cryogenic Fuel technology, AIAA 9-, 3th Aerospace Sciences Meeting and Exhibit, January 199. Smalley, A. J., Hollingsworth, J. R., Habets, G.L.G.M., and Kimmel, H. E., September 199, Rotor Dynamic Analysis of a Submerged Turbine Generator Driven by Liquefied Natural Gas, Proceedings of the Fifth International Conference on Rotor Dynamics, IFToMM, Darmstadt University of Technology, Darmstadt, Germany, pp. 9-39. Madison, J., March, Torque Dependent Vibrations for a Vertical Cryogenic Turbine Generator, Proceedings of the th International Symposium on Transport Phenomena and Dynamics of Rotating Machinery, ISROMAC, Vol. I. Guo, S., and Tano, M., Analysis on the Coupled Vibration of Rotor and Casing of a Vertical Shaft Pump by Using the Substructure Synthesis Method, Proceedings of the th International Symposium on Transport Phenomena and Dynamics of Rotating Machinery, Vol. 1, 199. Harris, R. E., and Bates, L. C., Southwest Research Institute Technical Memorandum --, June 199.