Copyright 1983 by ASME A STIFF, LIGHTWEIGHT MACHINERY BASEPLATE ON THREE SPHERICAL SUPPORTS

Transcription

1 THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS 345 E. 47 St., New York, N.Y GT-146 The Society shall not be responsible for statements or opinions advanced in papers or in discussion at meetings of the Society or of Its Divisions or Sections, or printed in its publications. Discussion is printed only if the paper is published in an ASME Journal. Released for general publication upon presentation. Full credit should be given to ASME, the Technical Division, and the authonsf. Papers are available from ASME for nine months after the meeting. Printed in USA. Copyright 1983 by ASME A STIFF, LIGHTWEIGHT MACHINERY BASEPLATE ON THREE SPHERICAL SUPPORTS Trevor Potter Principal Engineer Ingersoll-Rand Company Limited Wythenshawe, England ABSTRACT A study was carried out to develop a lightweight machinery baseplate for offshore applications. To prevent platform movement from affecting the machinery alignment, the baseplate was mounted on three bearings. Inherent stiffness, with significant weight saving over a conventional baseplate, was achieved by the use of a torque tube as the primary structural component. INTRODUCTION One of the problems resulting from the use of offshore platforms for oil and gas production has been the effect of platform distortion on the alignment of installed machinery. This distortion results from changing loads on the supporting structure due to re-positioning of the drilling rig, as well as environmental and/or thermal effects. Although this distortion is known to exist, its extent is not fully documented. Although flexible couplings have been provided to compensate for misalignment, the offshore sector of the oil and gas industry recognizes the need to eradicate this problem through preventive measures. This sector is becoming increasingly aware also of the effect of equipment weight on platform cost. Experience has shown that platform structures have to be overdesigned or strengthened because original equipment manufacturers have significantly underestimated their products weights. The requirements for the baseplate design for gas turbine driven a.c. generator packages for an offshore platform were: - The baseplate had to be mounted on three supports to minimise misalignment in the equipment train caused by platform distortion. - An acoustic enclosure, complete with exhaust silencer, ventilation and associated systems was required to cover the gas generator, power turbine and gearbox. - The mineral and synthetic lube oil systems (including storage tanks), the distillate and gas fuel systems had to be mounted on the baseplate. These three requirements are not uncommon, but a fourth requirement for minimum weight is a new departure and marks the introduction of a rigorous weight limit being placed on the total equipment package. Two concessions were made: that neither the gas generator air intake system, nor the lubrication oil cooler system need to be mounted on the basplate. CONCEPT DEVELOPMENT Conventional Design Traditionally, an equipment train consisting of a gas generator, power turbine, gearbox, a.c. generator or compressor would be supported by a baseplate constructed from two longitudinal 'I' beams, one along either side, with transverse beams and shorter longitudinal beams positioned to support the main equipment pieces. The 'I' beams would be approximately 15,000mm in length and approximately 760mm deep. The baseplate itself would be approximately 3500mm wide and, when installed, be supported along the underside of each longitudinal side beam at approximately 1500mm centres. The size of the longitudinal side beams would be dictated not necessarily by the weight of the equipment on the baseplate and associated operating loads, but more probably by the allowable baseplate deflection when lifted. The weight of such a traditional baseplate, including lubricating oil reservoirs, sized and constructed in accordance with API standards, would be of the order of 32,000 kilograms. Fig. 1 shows a simplified baseplate of this type. Modified Conventional Design When considering the requirements for this new baseplate design, it was decided initially to utilize a baseplate of traditional design but supported at three positions only. The dispositions of the three supports were; one under the gas generator on the baseplate longitudinal centre line and one either side of the baseplate under the a.c. generator. One essential requirement for power generating machinery is that it has to be capable of withstanding

2 tube above the single bearing. The final concept is illustrated in Figure 2. STRUCTURAL ANALYSIS 760mm FIG 1 'TRADITIONAL' BASEPLATE sudden short circuit loads four and a half times greater than maximum normal operating torque. A design evaluation conducted by a firm of consulting engineers revealed that a traditional baseplate, when mounted at three positions, would have adequate bending strength, but insufficient torsional resistance. Therefore, the baseplate would have to be further modified by the addition of a torque resistant system within the existing framework. It was concluded that a more efficient use of the steelwork was required to produce a baseplate which had to be: - Supported at three points in such a manner as to prevent normal platform deflections resulting in distortion of the structure and consequent machinery misalignment. Inherently stiff to withstand the static and operating loads without being overstressed, or causing unacceptable machinery alignment changes when supported at three positions only. Free from resonance excited by machinery operation. Fabricated from standard structural steels. Capable of containing the necessary quantities of machinery lubricating oils. The lightest structure practicable, commensurate with the satisfactory operation of the entire generating system when installed off-shore. Final Design Concept Having established the limitations of the designs for the modified traditional baseplates, new designs were prepared. These were based upon a single longitudinal torque resisting member as the primary load carrying structure, supporting rigidly fixed transverse sections to form a machinery platform. The sides of the platform were formed by members connecting the ends of the transverse members. The major transverse member supported two of the bearings at its extremities. The third bearing was connected to the torque resisting member on the baseplate centre-line. Various profiles of the torque resisting member were considered before the design which utilised a longitudinal tube was adopted. The final solution not only retained the rigidity of the torque tube connections but also incorporated rigid connections between the transverse members and the side members. This reduced the weight significantly but slightly increased the fabrication costs. The overall height of the machinery package, when mounted on its three supports was kept as low as practicable by reducing the diameter of the torque The critical design parameter for the machinery support arrangement was that of permissible misalignment of the couplings. This was measured from the true datum established from an alignment operation designed to bring all shafts to their correct position in all planes with the machinery package in stand-by mode. (That is with full oil tanks, all design superloads but with all equipment inoperative). Deflections of the structure under this design condition were to be within normal structural tolerances because these displacements have no effect upon the subsequent alignment of the shafts. Generally, stresses were to be based on those given in A.I.S.C. Specification for the Design, Fabrication and Erection of Structural Steel for Buildings. However, the intention was to use steel to British Standard and not to A.S.T.M. Specification, since British Standard material is more readily available in the United Kingdom. Certain stress limits were therefore based on the yield stress for material to B.S The loading data used for the structural analysis included the dead weight, superloads, operating loads, and the effect of the plenum chamber depression produced when the gas generator is operating. Thermal effects due to operating temperature gradients were also covered. In addition to the static analysis a dynamic analysis was necessary to compare the natural frequencies of the structure with the operating speeds of the equipment to ensure that vibration did not result from resonance. The following design cases were statically analysed: 1. The machinery package in stand-by mode, that is, with full oil tanks, all design superloads but with all equipment inoperative. 2. The normal operating condition, that is, as in design case (1) but with the machinery operating at full load. 3. The normal operating condition as described in (2), with the addition of loads due to a short circuit fault condition. 4. The normal operating condition (2) with additional loads due to a power turbine seizure. The results of this static analysis included loads and stresses in all members and displacements in all planes. These displacements were used to calculate the change of machinery alignment between Design Case (1), that is, with the machinery package in stand-by mode, and each of the other design cases. The resulting shaft displacements are shown in Figure 3. The resulting coupling misalignment, compared with that permissible, is shown in Figure 4. The displacements shown in Figures 3 and 4 are for the situation described in Design Case (3) only. The figures show that, for all operating conditions, the coupling misalignments were well within the limits specified by the manufacturers. The stresses and deflections for the baseplate itself proved to be more critical than the coupling misalignments, although well within the allowable 2

3 Rigid Torque Resisting Member mm 0/A Location Key Spherical Bearing Assembly Incorporating a Sliding Joint 1200mm 250mm 1104P END ELEVATION Spherical Bearing Assemblies 3500mm 0/A FIG 2 FINAL CONCEPT levels for the materials used. The corresponding vertical misalignment between the power turbine and the gearbox was 0.4mm. This, when added to the previously calculated misalignment due to normal operating and fault conditions, was also well within the couplings' limitations. The effect of the possible temperature gradient (26 C) from top to bottom of the torque tube, was a relative upward deflection of 4mm at the mid-point of the length of the tube. The corresponding bending stress was approximately 35 N/mm 2. These effects, when added to the uniform temperature stresses for Design Cases (1) and (4), were within the allowable levels. A comparison between the actual and permitted stresses and deflections is given in Table 1. The maximum deflection is 6.1mm and occurs approximately mid way between the front and rear bearings. Maximum stress also occurs at this point. Some refinement of the member sizing may have been possible to further reduce weight but the production programme necessitated freezing the design at that stage. Even so, the weight for the baseplate complete with oil tanks was only 18,000Kgs compared with 32,000Kgs for a 'traditional' baseplate referred to earlier. The analysis output for the dynamic behaviour of the baseplate produced a list in ascending order of 52 natural frequencies with relative displacement at each mode considered. Because of the form of vibration, only vertical displacements were significant and only these were tabulated. The design philosophy was based on resonance avoidance rather than resonance effect, so further analysis was not carried out to establish the dynamic response of the structure for the various modes. The results showed that the configuration of the baseplate on three supports gave rise to two possible fundamental forms of vibration : torsional and flexural. Higher frequency levels corresponded to harmonics and combinations of the two fundamental modes. The eight lowest frequencies were considered to cover a sufficiently high order of harmonics for comparison with possible excitation frequencies. TABLE 1 Stresses and Deflections Max. Vertical Bending Stress in Main Longitudinal Torque Tube STRESS COMPRESSIVE BENDING STRESS Cold Operating Case Thermal Effect (approx) Total CO-EXISTENT TORSIONAL SHEAR STRESS EQUIVALENT COMBINED STRESS ACTUAL MAX ALLOWABLE 58 N/mm2 35 N/mm2 93 N/mm2 165 N/mm2 22 N/mm2 100 N/mm2 101 N/mm2 230 N/mm2 3

4 r MM PROJECTED END OF SHAFT d=angular DISPLACEMENT HORIZ. DISPLACEMENT PLANE (M M) VERT. PLANE Position A Position B Position C Position ' DISPLACEMENT (MM) ANGULAR DISPLACEMENT (Ef) MORK PLANE VERT PLANE HORIZ PLANE VERT PLANE ACTUAL PERMISS ACTUAL PERMISS ACTUAL PERMISS ACTUAL PERMISS NEGLIGIBLE FIG 3 ABSOLUTE DISPLACEMENT OF THE POWER TURBINE AND GEARBOX INPUT SHAFT ENDS Modes with higher frequencies, while theoretically capable of excitation, were unlikely to occur with significant amplitudes. The mode shapes corresponding to these eight lowest frequencies are shown in Figure 5. The most likely cause of excitation is the out of balance force for the rotating equipment. Table 2 lists the rotation speeds, with the corresponding frequency, of the various shafts under both idling and operating conditions. This table also gives teeth meshing frequency of the gearbox. TABLE 2 Excitation Frequencies Source Alternator with P.T. Idling at 3,000 rpm Alternator Synchronised Rotational Speed r.p.m Corresponding Frequency HZ FIG 4 ACUAL AND PERMITTED COUPLING MISALIGNMENT In order to avoid the 17.5 Hz natural frequency of the structure the analysis showed that the idling speed for the power turbine should not be less than 3000 rpm, so that the corresponding a.c. generator speed was not less than 1137 rpm. This did not present a problem for, when the equipment was commissioned, the idling speed of the power turbine could be set above the required minimum by slightly increasing the idling speed of the gas generator. Comparison of the natural frequencies shown in Fig.5 with the excitation frequencies shown in Table 2 shows that during run up of the machinery to idling speed, all the shaft speeds correspond at some time to the first eight natural frequencies. Once idling speed has been reached, all the shafts would be running at speeds greater than these frequencies. The gear meshing frequency was so high that it did not pose a problem. Due to the accuracy of the balance procedures used for this type of rotating machinery, the degree of out of balance would be very small. Therefore, passing through the natural frequencies during run-up and run-down was not likely to produce vibration levels of any significance. This was borne out during works testing of the equipment (albeit without the a.c. generator) at Ingersoll-Rand's Wythenshawe England plant in June P.T. with Avon Idling Avon Idling P.T. with Alternator Synchronised Avon Running Range Gearbox Teeth Meshing to to SUPPORT BEARINGS The baseplate was designed to be mounted on three bearings as shown in Fig.2, each of which was to permit rotation about both longitudinal and transverse axes. The type of bearing selected to meet the load carrying and displacement requirements was manufactured by Rose Bearings. The Rose bearing, which consists of an aluminium bronze ball contained within a chromium steel housing, was mounted in a bearing assembly as shown in Fig.6. The design of the bearing assembly ensures that the only effect on the baseplate from rotational and vertical movements of the supporting structure deck is 4

5 MODE 1 MODE 2 f= ,,, MODE 4 f=10 3H3 MODE 5 1=12-4H, MODE 6 f=12.9 H3 MODE 7 (-158 H 5 MODE - 8 f=17.5h-3 A sliding joint was devised which provided transverse location and allowed longitudinal expansion to take place. The two rear bearings placed at the heavier end of the equipment train were connected to the larger diameter tube through a box beam, again shaped so as to keep the overall height, with the bearings fitted, to a minimum. The reduction in tube diameter beyond this box beam was done purely as a weight saving exercise but it is debatable whether the small reduction was justified in view of the added fabrication complication. It is envisaged that future baseplate designs would utilise a continuation of the larger tube. Originally it had been intended to utilise the large volume of the torque tube as the mineral lubrication oil reservoir but API Standards, to which the lubrication system had to be designed, dictated that large manways had to be provided for access. It was not practicable to incorporate the requisite manways in the tube and retain the necessary torque resisting characteristics. However, one advantage of the reservoir as positioned was that the drains from the equipment utilising mineral oil for lubrication could be piped directly through short pipe runs into the reservoir, and thus the usual large diameter drain pipe runs into system was eliminated. The completed baseplate and driving equipment is shown in Fig.7. Xji FIG 5 MODE SHAPES AND FREQUENCIES stiff body movement, which causes no distortion of the baseplate. The two rear bearings were arranged in line to allow the degree of lateral movement required for thermal effects by permitting the ball to slid along the shaft. A laterial stop was provided on the longitudinal centre line of the transverse bearing beam to locate the baseplate. All longitudinal forces are resisted by the rear bearing. Longitudinal movement of the baseplate due to thermal effects was allowed for by the provision of a longitudinal sliding joint at the single front bearing position. DETAILING The transition from the larger diameter tube to the smaller diameter tube was necessary to enable the single front bearing to be placed on the centreline of the tube and to keep the overall height as low as possible. This was achieved by the use of high tensile material, and is the only use of non-standard steel. The method of achieving strength at the front bearing location dictated that the front bearing assembly had to be positioned with its shaft transverse to the tube centreline. With the shaft in this position it precluded the use of the inherent freedom of the Rose bearing to slide along its shaft to allow for longitudinal expansion of the baseplate. CONCLUSIONS The baseplate was designed to support a gas turbine driven a.c. generator package for an offshore application. The requirement that normal platform deflections should not result in distortion of the baseplate, with consequent machinery misalignment, has been achieved by mounting the baseplate on three spherical bearings so positioned that regardless of their relative movement due to platform deflections, they remain in a plane. The effect of baseplate longitudinal expansion has been allowed for in the design of the single front bearing location. The stiffness, both torsional and flexural, necessary to withstand the worst operating conditions without producing unacceptable machinery alignment changes, has been achieved by the use of a torque tube and main transverse box beam linking the three bearing positions. It is this basic configuration which also gives the majority of the weight reduction over a 'traditional' longitudinal and lateral beam design. The configuration, with its inherent torsional and flexural stiffness, has enabled the basic design criteria of maintaining machinery alignment, within defined permissible limits during operation, to be achieved with a considerable margin. The deflection and stress requirements for the baseplate have been met successfully. The dynamic analysis has produced a list of natural baseplate frequencies which shows that the baseplate will not be excited by operation of the machinery in the package. The baseplate has certain inherent advantages over the 'traditional' designs in addition to its lighter weight and three spherical supports. These are: - It has the property of machinery alignment being maintained if the baseplate is picked up and placed in a different location on the three supports. It provides a flat working surface for installing 5

6 FIG 6 TYPICAL BEARING ASSEMBLY and operating equipment. The space below the flat working surface enables pipework, cables and other equipment to be mounted with easy access from either side of the baseplate. It has the capability of a single point lift. It is suitable for all types of equipment required to be 'skid' mounted either on-shore or off-shore. The design, although based on fundamental principles, conventional components and standard steelwork practices, is so different from any known previous design that worldwide patent applications have been filed. ACKNOWLEDGEMENTS The author gratefully acknowledges the assistance received from Zenith Cape Ltd. Consulting Engineers. Lincoln, England; and Anthony Cropps, the Assistant Chief Draughtsman at Ingersoll-Rand Wythenshawe. FIG 7 COMPLETED BASEPLATE AND DRIVING EQUIPMENT 6