TRIUMF
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TRIUMF brown reports TRI-68-1
Dilworth, Secord, Meagher and Associates Limited
CONCEPTUAL DESIGN STUDY FOR THE SUPPORT STRUCTURES, ELEVATING JACKS AND LEVELING JACKS FOR A 500 MEY H- CYCLOTRON D'ILWORTH, SECORD, MEAGHER AND ASSOCIATES LIMITED UNIVERSITY OF ALBERTA SIMON FRASER UNIVERSITY UNIVERSITY OF VICTORIA U IVERSITY OF BRITISH COLUMBIA I I I -I-I I I I I I I I I I I I I I I TRI-UNIVERSITY MESON FACILITY CONCEPTUAL DESIGN STUDY FOR THE SUPPORT STRUCTURES, ELEVATING JACKS AND LEVELING JACKS FOR A 500 MEV H- CYCLOTRON by TRI-68-1 DILWORTH, SECORD, MFAGHER AND ASSOCIATES LIMITED Postal Address: TRIUMF, University of British Columbia, Vancouver 8, B.C. Canada July 1968 I I I -I-I I I I I I I I I I I I I I I ,... CONTRIBUTORS The following professional staff members of Dilworth, Secord, Meagher And Associates Limited contributed to this report: U. Kiesners W.J.F. Bradley M.A. Fisher J.D. Stringer L.J.P. Tillson I I -I-I I I I I I I I I I I I I I I SUMMARY Studies of concept feasibility with alternative cost oriented.design consideration for the supporting structures indicate that the basic con-cept of the cyclotron support structure, vacuum tank and magnet gap as established by TRIUMF is feasible. The recamnended primary top structure would consist of straight I-section beams welded radially to a stiff centre ring. Secondary structure would be bolted to the primary in order that the maximum crane lift be limited to approximately 100 tons. The top and bottan of the vacuum tank would be 7/8 inches thick, supported by 1-1/2 inch diameter tie rods at a spadng not greater than 33 inches. Leveling of the magnet cores would be accanplished by a system con-sisting of hydraulic actuators for raising, used in conjunction with mechanically motor operated wedges for precise height setting. Alte~ natively, motor operated Acme screw jacks of 150 tons capacity operating through simple lever mechanisms should be considered further. To elevate the top structure and upper magnet cores, a system util-izing 150 ton Acme screw jacks synchronized to control vertical setting within 0 . 02 5 inches is recarmended. Protective devices to prevent crushing of secondary coils between vacuum tank and magnet faces have been introduced. It is indicated that these stops will permit greater structural deflections with attendant decrease in beam depth and as such warrant further study. The fabricated canponents embraceo in this report are well within the manufactur•ing capabilities of several local firms. I I TABLE OF CONTENTS P~e No. I - 1. INTRODUCTION 1 I- 2. SCOPE 1 3. DESIGN CRITERIA 2 I 3.1 Space Restrictions in Magnet Gap 2 I 3.2 Space Restrictions in Vacuum Tank 3 3.3 Leveling Jack System 3 I 3.4 Elevating Jack System 3 4. INVESTIGATION F SUPPORT STRUCTURE AND TANK 3 I 4.1 Arch Concepts 3 4.1.1 Analysis of Arch 4 I 4.2 Straight Beam Concept 7 ,, I 4.2.1 Analysis of Straight Beam 7 4.3 Canparison of Arch Versus Straight Beam 7 I 4.3.1 Detailed Canparison 8 4.4 Secondary Structure 10 I 4.5 Tie Rods 10 I 4.6 l.Dwer Supporting Structure 11 4.7 Centre Post 11 I 4.8 Vacuum Tank Concepts 12 4.8.1 Pressure Effects on Tank 12 I 4.8.2 Thermal Effects on Tank 13 4.9 Protective Devices 13 I 4.9.1 Selection of Suitable Stops 14 4.9.2 Analysis of Structure for Increased I Deflection 15 5. MANUFACTURING CONSIDERATIONS 17 I I I I I. I-I I I I I I I I I I I I I I I 6. 7. 8. Page No. 5.1 Vacuum Tank 17 5.1.1 Fabrication Tolerances 17 5.1.2 Mill Tolerances 17 5.1.3 Overall Manufacturing Tolerances 18 5.2 Support Structure 18 INVESTIGATION FJACKS 18 6.1 General 18 6.1.1 Ball Screw Jacks 18 6 .1. 2 Acme Screw Jacks 19 6.1.3 Motor Driven Wedges 19 6.1.4 Hydraulic Actuators 19 6.2 Alternative Jack Ar.r>angements 20 6.2.1 Magnet Leveling Jack Systens 21 6.2.2 Elevating Jack System 21 6.3 Control Systens 21 6.3.1 Control Systens for Leveling Jacks 22 6.3.2 Control Systems for Elevating Jacks 22 6.4 Operating Environment 24 6.5 Evaluation of Alternative Jacking Systems 25 6.5.1 General 25 6.5.2 Magnet Leveling Jack System 25 6.5.3 Elevating Jack System 26 COST ESTil1ATES 26 7.1 Supporting Structures 26 7.2 Jacks and Controls 27 7.2.1 General 27 7.2.2 Magnet Leveling Jack Systems 28 7.2.3 Elevating Jack Systems 28 RECOMMENDATIONS 29 8.1 Supporting Structures 29 .... 8.2 Jack Systems 8.2.1 8.2.2 Magnet Leveling Jack Systems Elevating Jack Systems 9 . FIGURES AND DRAw.m3S Fig. la System (a) Fig. lb System (b) Fig. le System (c) Fig. 2 Maxinn.un Deflection of Arch Versus Arch Rise Fig, 3 Maxinn.un Deflection of Arch Versus Ring krea Fig. 4 Translation of Outer-Arch Support Versus Restraining Ring Area Fig. 5 Translation of nuter Arch Support Versus Arch Rise Fig. 6 Schematic Elevations of Arch Structure Page No. 30 30 30 Fig. 7 Manent of Inertia Values Versus Deflection of Straight Beam Fig. 8 Schema.tic of Straight Beam Fig. 9 Total Structural Weight Versus Arch Rise Fig. 10 Double Arch Concept - 48" Df'ep Section Fig. 11 Schematic I.Dading Diagram of Lower Concrete Supporting Structure Fig. 12 Closed Loop Control System for leveling Jacks Fig. 13 Closed l.oop Control System for Leveling Jacks Fig. 14 Standard Selsyn Control System for Elevating Jacks Fig. 15 Closed Loop Control System for Elevating Jacks Drwg. 60000 Layout of Tie Rods Drwg. 60001 Support Structures Straight Beam Concept I)n...tg. 60002 Centre Post I I 1-I-I I I I I I I I I I I I I I I .. 1. 2. 1. INTRODUCTION The studies herein are in accordance with requirements of Simon Fraser University purchase order number 25484 dated March 12, 1968. The design criteria and scope of work are stipulated in a letter dated December 13, 1967 from J.J. Burgerjon of TRIUMF to F.G. Pearce of Dilworth, Secord, Meagher and Associates Limited. By maintaining close liaison with the TRIUMF group throughout this assignment, it has been possible to introduce additional beneficial features to the design, to modify details of the original concept where indicated and to correlate the work with other features of the cyclotron out-side the scope of this study, thus avoiding possible conflict with other requirements of the overall design. SCOPE To establish the feasibility of the TRIUMF concept, it was required to satisfy the following conditions: ( a) Develop an overhead structure capable of withstanding the forces due to atmospheric load, magnetic load and structure dead weight within allowable deflections stipulated by TRIUMF. (b) (c) (d) (e) (f) Utilizing the same structure, to safely elevate the structure together with the top halves of the magnet cores and the top of the vacuum tank when required for maintenance purposes. Investigate the deflections of the vacuum tank covers under vacutpn conditions and determine the spacing of suitably dimensioned tie rods to restrict this de-flection within the clearance limits specified. Analyze conceptually a concrete base to deflect elastically not more than the upper structure when acted on by the upward forces exerted by the lower tie rods under vacuum conditions. Investigate the possibility of providing a centre post of the geanetrical configuration required by TRIUMF which would accomnodate safely the high stress developed in this member. Provide a system of jacks to level and support each magnet sector. I I I I-I I I I I I I I I I I I I I I 3, (g) (h) (i) Provide a system of jacks to elevate the upper support structure and magnets for servicing of the cyclotron. Develop a controls system for each jacking system for accurately controlled actuation of leveling jacks and synchronization of elevating jacks, 2. Investigate the feasibility of manufacturing the support structures and vacuum tank, with special emphasis on manufacturing tolerances required to maintain the dimen-sions of the vacuum tank within clearances stipulated. FlLYTther investigation during the course of the assignment indicated: (a) (b) DESIGN CRITERIA The desirability of providing safety devices to prevent crushing of secondary coils against the magnet faces under conditions of vented tank or structure elevating. Requirements for a study of the buckling effects on the vacuUIJl tank when subjected to a "bake out" temperature of 200°c. 3.1 Space Restrictions in Magnet Gap The 20 inch gap between the upper magnet face and the lower magnet face is a fixed dimension and may not be increased without incurring a severe penalty to magnet design. Within this gap, a vacuum tank is suspended by means of tie rods connected to the upper and lower support structures. During erection, the top and bottan surfaces of the tank are adjusted under vacuum load so that the inside plate-to-plate dimension is 16.25 inches and the faces are parallel. This dimension is established by the resonators within the tank. Using 7/8 inch plate construction for tank lid and bottan, and 3/4 inch diameter coils attached externally to the plates, the remaining space between the coils and magnet faces, when the tank is under vacuum, is 1/4 inch on each side of the tank (less manufacturing tolerances). As the tank is vented and the atmos-pheric load is released, the structure supporting the tank will spring back to its unloaded position. If the springback deflection is greater than the 1/4 inch available (less manufacturing tolerance) then damage to the coils could result by crushing against the magnet faces. In addition, this same system must be able to withstand loads due to weight and magnet forces when operating, and to elevate the lid and upper magnet sectors safely for maintenance when required without damage to the coils. I I I I-I I I I I I I I I I I I I I I 4. 3. 3.2 Space Restrictions in Vacuum Tank 'Ihe resonators occupy a large portion of the space within the vacuum tank and are suspended fran the tank lid and bottan at tie rod attach-ment points. As presently envisaged, these resonators are positioned with only 1/8 inch clearance between their surfaces adjacent to the tank top and bottan. In addition to the requirements of .3.1 above, the summation of deflections of the tank under vacuum conditions plus manufacturing tolerances must be accancxiated within these.clearances in order that the overall gap of 2 0 inches between magnet faces may be maintained. 3.3 Leveling Jack System 'Ihe magnet leveling jacks must be capable of supporting and leveling the six magnet core assemblies, each supported on two leveling jacks and a carmon centre post. The total load of 12 jacks is 3100 .tons. Two inches jack stroke is required for original alignment of the mag-net cores during installation and 0.25 inches remotely controlled adjustment is required for fine alignment during carrmissioning and further adjustment at a later date if any misalignment has taken place. Each of the 12 jacks must be independently operated and controlled to obtain position accuracy within plus or minus 0.010 inches. Raising or lowering speed of the jacks is not critical, and naninal speed of 0. 4 inches per minute has been chosen. 3.4 Elevat~ Jack System The elevating jacks must be capable of elevating the top structure, upper magnet cores and top cover of the vacuum tank a distance of 36 inches above the nonnal operating condition for access into the vacuum tank. The total dead weight to be elevated is 1740 tons and this is distributed equally on the 12 elevated jacks. All the jacks must be synchronized, and gain or lag of any individual jack must not exceed a limit which would cause unacceptable structural defonnation .. of the elevated assembly. This limit must be established during the course of this study. The required speed of elevation is 2 inches per minute. INVEsrIGATION OF SUPPORT srRUCTURES AND TANK 4.1 Arch Concepts The arch concept has been suggested as a means of obtaining a lighter weight for the overhead structure supporting the vacuum tank. The requirement of any support structure is to transfer loads applied by the vacuum tank and tie rcxis (skyhooks), out to the outer support columns for transfer to the ground. These loads can occur under operating conditions when the vacuum tank is evacuated, and under elevating conditions when the magnets and tank cover are lifted for servicing. I I I .. I-I I I I I I I I I I I I I I I 4. Arch structures of various configurations are possible in this appli-cation. Elevations of these are shown schematically in Figures la, b and c. System 'a' is referred to as the double arch concept and is canposed of two separate arch systems. The one having the short span carries the operating loads (principally atmospheric pressure. on vacuum tank) and the long span arches carry loads due to elevating the overhead structure. System 'b' , the single arch concept, has a radial pattern of single long span arches across the diameter of the magnet layout. The maximum rise occurs at the centre where the arches meet. The single, short span system 'c' is made up of arches designed for the operating loads and reinforced by doubler plates and cross brac-ing to handle the elevating loading. The cross bracing, made up of horizontal canpression struts, runs radially inward between the high-est points of the opposite arches. Additional stiffening required for the outer half of the arches copes with the shears and manents arising fran the elevating loads. Both system (b) and (c) have structural limitations when canpared to the double arch concept. System (b) is at a disadvantage because the attainable arch rise to span ratio permitted for the operating case is significantly less than that allowed in the double arch system. A shallow arch when subject to the loading and geometrical constraints imposed by the cyclotron has a greater weight for a given deflection and hence is less economical structurally. This is known fran the basic properties of arches and is confinned by the analytical study. Calculations show that the total weight of system (c) due to addition of intersecting compression members and necessary reinforcement of the outer ends of the arches exceeds the weight of separate long span arches designed solely to carry the elevating loads. Considering this, the double arch concept of system (a) idealizes the arch design and is used in carrying out the detailed analysis. The double arch concept can be presented as an optimum .:rrich design when comparing with another type of support structure, i.e. straight beam. 4.1.1. Analysis of Arch The arch structure is composed basically of radial arches joined at their inner ends and linked together at their outer ends by a peri-pheral ring composed of straight segments of equal length. The design optimization was performed by analyzing several different arches having various rises and combined with restraining rings of various stiffnesses. The calculations were performed for the case of operating loads applied to the structure and arches of the following geometries were considered: 5!-•tl - ,...(-,...ti;, rA-f-"':. .....---A, (J,. c-t.t..e: Mt;;. ry } l;.:.t.:,e 7 ;t.... .,,.n.# .f- - ~I( - c.ll(( . .(,... -_... .. t;;. ...,. tt.,..t ---, 11:--......ts-- ... ~K . tAU 4'.. ,~ ~•A ~l- £,. '->=•-. -f:' h- _,..__ ~,.A-I I 1-I-I I I I I I I I I I I I I I I IJ(..,.,, r; r -Arch Span 37811 378" 37811 Arch Rise 60" 9011 120" 5. For each of these gecmetries, peripheral restraining rings of var:ious cross sectional areas were considered. The areas used were 40 in.2, 80 in. 2, and 120 in. 2 and the peripheral ring was assumed to carry direct tension only due to operating loads on the arches. As a result, nine redundant structures were analyzed. Resulting bending moments, endloads, shears, and deflections at seven points of interest on the arches and restraining ring were found for each of the nine geometries. To find these values , the strai.i"1 eneK§Y method was used. 111e redun-dant systems were first solved forte unknown redunclants and then these redundants were applied to the structural system to find the true loads and the true structural deflections. The complete results were sumnarized by plotting the results for each geanetry on graphs re-lating the deflection of the arch to: 1. the span of the arch 2. the rise of the arch 3. the stiffness ( cross sectional area) of the peripheral restraining ring Interpolations were made fran graphs to estimate the effect of inter-mediate geometries. An example of curves is given. Figure 2 shows the relationship between maximum vertical deflection of arch and the arch rise for structures having various peripheral restrainjng ring areas. Figure 3 relates maximum deflection to the area of the peri-pheral restraining ring for various arch rises. All these calculations are based n the same assumed arch cross section. Choice of structural section sizes was made to ensure practical end connections, ·to avoid longwave flange buckling, local instability of flanges and webs and web crippling. For fabricating simplicity on the end connections, the practical arch depth is limited to about 30 inches. later, in the weight study, various arch depths were studied up to the limit of available standard sections. The peripheral ring which restrains the outer ends of the twelve arches, acts as an elastic support due to its own elasticity under tensile load-ing. This elastic deflection permits the outer ends of the arches to translate horizontally. Figures 4 and 5 illustrate this deflection as a function of restraining ring area and arch rise respectively. Such a tran lation reduces the effectiveness of the arch in carrying vertical loads . The maximum arch deflection under the vacuum tank loads increases as the horizontal restraint reduces. This horizontal restraint reduces as the area of cross section of the peripheral ring is lowered. I I 1-I-I I I I I I I I I I I I I I I 6. Figure 3 which plots maximum deflection against peripheral ring area illustrates this. It can be seen fran this graph, that the increase of ring area has a non-linear effect on the maximtm arch deflection. At increasing ring areas, the rate of reduction of maximum arch deflection decreases. The graph shows that at peripheral ring areas above 100 to 120 square inches, very large increases of ring area are necessary to appreciably reduce the maximtm arch deflection. Because the peripheral ring provides the added fun~tion of carrying a part of the upper magnet weight when elevating, its function as_a beam must also be considered. A restraining ring area of approxi-mately 90 square inches is optimum for both considerations. The other variable, arch rise, indicates the effect of the arch shape on the maximum deflection under a given loading. This arch shape, represented by a rise or vertical arch height for a given span length, applies to circular arches only in this study. The results indicate that increasing the arch rise consistently reduces the maximum arch deflection when loading, span and section properties do not change. This occurs for the various values of peripheral ring areas used in the study. For example, the maximum arch deflection reduces fran 0.2 inches to 0.1 inches when the arch rise is increased from 60 inches to 108 inches. The peripheral ring area for that example equals 80 square inches. The effect of arch rise on max-imum deflection is non-linear and the rate of change of deflection decreases as the arch rise increases. For the present geometry and loading, an optimum arch rise of approximately 105 inches is indicated. AfxJve this value, the reduction in maximum deflection that results is almost negligible. It can be concluded that the optimum arch geometry for carrying the operating loads would be a canbination of twelve arches having rises of 105 inches that are restrained by a peripheral ring of 90 square inches cross sectional area. A change in the section properties of the arch will have no effect on the general conclusions stated above. Deeper sections would, however, give reduced maximum deflections for the arch. The problem introduced by deep sections when connecting the elevating and operating arch systems together and restraining the ends, limited their effectiveness in reducing deflections. Because of these practical reasons, the maximum deflection of the arches cannot be reduced much below 0.1 inches. The part of the double arch structure carrying the loads due to ele-vating is a long span circular arch bridging the full diameter of the cyclotron structure. As seen in Figure 6 , the optimum rise of this elevating arch is al:x:>ut 200 inches in order to avoid interference with the shorter span operating arches and to minimize the total structural height. A deflection analysis and stress check on an 18WF 7 0 section for this application indicates that a maximum deflection of 1/8 inches would occur with maximum bending stress of about 12,000 psi. I I 1-1· I I I I I I I I I I I I I I I 4.2 Straight Beam Concept The geanetry and section properties of a straight primary beam were developed and optimized to provide a definite maximum vertical de-flection under the operating loads of atmospheric pressure on the vacuum tank. Several beams of the 'I' beam type having different section properties and depths were chosen and analyzed for deflec-tion. Deflection due to both bending and shear deformation were considered, and the results of these calculations are shown in Figure 7. 4.2.1 Analysis of Straight Beam The method used was of the strain-energy type including shear in the analysis. The beam configuration is assumed to be a propped cantilever beam built in at the inner end by attachment to a stiff ring of small diameter. This ring carTies the inner end of all twelve beams. The outer ends of the beams are supported on jacks used for elevating the canplete structure. The operating load of 7. 370 kips is distributed along the beam, and deflections were cal-culated at 12 node points. Figure 8 illustrates the beam configura-tion and loading. Figure 7 presents the manent-of-inertia values versus the maximum deflections for seven beams of various depths, web thicJcness, flange widths and moment-of-inertia values. The graph includes a 95% confidence band. This indicates the region within which a maximum total beam deflection would lie for any particular ~anent-of-inertia value with 95% confidence of being right. The est.:i.mate of the band width is based only on beams tested and more beam calculations would almost certainly narrow it. It can be concluded that a manent-of-inertia value of 150,000 in. 4 is required to limit maximum deflection of the straight beam to 0.050 inches under the operating loading condition. Under elevating conditions the maximum deflection is 3/8 inches with a maximum bending stress of approx.:i.mately 8000 psi. This increased deflection over that of the operating condition will reduce the gap between vacuum tank cover and magnet cores during elevating conditions. However, the use of mechanical stops as des-cribed in Section 4.9 and illustrated on Drawing 60001 will prevent the tank cover and secondary coils fran possible damage caused by this increased deflection. 4.3 _£an~ison of Arch Versus Straight Beam The two concepts are canpared principally on the basis of overall weight and est.:i.mated cost. The desire to handle the overhead structure by travelling crane and to reduce structure and building costs encourages the use of a light-weight, yet canparatively simple structure of canpact vertical dimensions. Initial considerations indicated that the requirements were not likely to be found in one type of structure. However, results of a weight analysis for opti-mum designs of both arch and straight beam concepts indicates that I I I-I· I I I I I I I I I I I I I I I 8. no weight saving would ·be ·gained by using a more structurally jnvolved, if more elegant arch concept. In addition, the straight beam concept has the advantage of greater structural simplicity and reduced height with consequent reduced manufacturing costs and decreased building height. • 4.3.1 Detailed Canparison 'The canparison of the two concepts was made by finding the total weight of the overhead support structure for each including'the primary struc-ture of radial members and restraining ring, the secondary structure of tangential members, and the tie rods and fittings. The two concepts·· considered were arch systems having various rises and a radial straight beam optimized to reduce weight. For canparison purposes, the maximum vertical deflection·under load for both concepts was taken as 0.1 inches for the primary radial members. The weight of·the·total overhead support structure was evaluated for the concepts when their designs met this criteria. The 0 .1 inches .. deflec-, tion was arbitrarily chosen because the geanetries and section proper-ties selected for practical considerations conveniently produced primary structure deflections of about this amount. Using Figures 2 and 3, the arch and restraining ring geanetries which give a maximum vertical deflection of 0.1 inches were selected. Using canputer calculations, the results of which are summarized in Figure 7, the straight beam geanetry and section properties necessary to provide 0 .1 inches maximum deflection were found. Fran this infonnation and estimates of the tie rod and secondary structure weights, a total esti-mated weight was found for each concept. Table 1 Structural Weights for Arch Concept Arch Rise (Inches) 90 105 120 Primary Structure Weight (Pounds) 294 000 274 000 264 000 Total Structure Weight (Pounds) 359 000 339 000 329 000 I I I I-I I I I I I I I I I I I I I I Table II Structural Weights for Straight Beam Concept Beam Depth (Inches) 45 60 Primary Structure Weight (Pounds) 327 000 197 000 Total Structure Weight (Pounds) 420 000 290 000 9. The valu-:s of total structure weight versus arch rise are plotted in Figure 9. This figure illustrates the ef feet of arch rise on the total weight of the arch concepts and also includes the weights of the straight beam concepts plotted on the ordinate for canparison. Arch structures in the above calculations have constant beam depths of 30 inches chosen from practical considerations. To investigate the effect of changing the section propertiefl, arches of different depths were chosen and their required properties to establish a maxi-mum deflection of 0.1 inches were found. Table 3 tabulates these findings. Table III Total Structural Weight Depth of Arch Member Arch Rise 36" 4211 48" (Inches) (Pounds) (Pounds) (Pounds) 90 325 000 302 000 278 000 105 302 000 279 000 254 000 120 290 000 268 000 241 000 The depth of the arch members has been investigated to the limit of commercially available fabricated sections. The section depth of an arch having the span and rise required for the operating loads is out of proportion to other dimensions as Figure 10 illustrates. End-connec-tion and fabrication problems are increased ccmpared to a straight beam. Figure 9 indicates that little weight saving can be gained by increas-ing the arch rise above 120 inches. The curve of total weight versus arch rise flattens out appreciably beyond that point and the weight reduction on total structure becomes negligible. The graph also shows I I I I-I I I I I I I I I I I I I I I 10. that use of the arch concept results in no appreciable saving of weight for the total structure when canpared with the weight of the straight beam concept. The largest contributing factor responsible for this result is the elastic flexibility of the peripheral restraining ring in the structure. This effect prevents the arches fran developing the full stiffness that geanetry and rigid supports would nornially allow. Figure 3 illustrates this. In conclusion, an arch of practical dimensions presents no weight saving over the straight beam concept, its fabrication and support is more difficult and costly and it requires more vertical space to accan-modate it. Regarding the straight beam structure, it is seen that increased beam de-pth greatly reduces structural weight. A weight reduction of one-third results fran increasing the straight beam depth fran 45 inches to 60 inches without changing the 0.1 inches deflection criterion. Calculations have also shown that the maximum beam deflection can be limited to 0.050 inches by increasing the beam depth from 60 inches to 90 inches and this can occur with no increase in total structural weight over that for the 60 inch deep beam. The total weight of over-head structure in that case is 290 kips. 4.4 Secondary Structure Composed of interconnecting beams between primary structure, the secondary structure is designed to deflect a maximum of not more than 0.050 inches under the atmospheric pressure load applied through the tie rods. This total deflection limit is achieved using channel type beams, gussetted to the primary radial beams at their ends, and pro-vided with tie rod fitting brackets along their length. Drawing No. 60001 illustrates the structure. 4.5 Tie Rods These members transfer the atmospheric pressure load of the operating case along their length in tension from the vacuum tank surface to the overhead primary and secondary structuY'e, or the concrete foundation. Spacing of tie rods is kept about equal at 33 inches between rods. This prevents over-stressing and over-deflection of the vacuum tank surf ace and maintains approximately equal load in the rods. A rod diameter of 1-1/2" ensures that tie rod elongation under load does not exceed 0.050 inches and stress level 10,000 psi. The outer ends of the upper magnet cores are suspended fran the over-head support structure when elevated. Steel rods of 10 inch diameter are adequate for this requirement and introduce a relative deflection between vacuum tank lid and upper magnet core of less than 0.003 inches at each rod location. I I 1-I-I I I I I I I I I I I I I I I 11. 4.6 Lower Supporting Structure A concrete foundation supports the cyclotron, its function being not only to provide a monolithic supporting 'raft' but also·to·furnish an anchor for the lower tie rods to restrain the vacuum tank bottan In this latter capacity, it must act as a restraining member with·a ~-imum deflection of 0.100 inches under load. Because this function of the base must eventually be integrated into-the foundation ·as a. -whole,. no detailed analysis was considered appropriate at this stage. How-ever, the following analysis demonstrates the feasibility of this method of support. • tion under the full area of the cyclotron. The pad carries the centre Y column and outer columns supporting the overhead structure, and. pro-vides tie-ddWri connections for the lower tie rods. The concrete pad is loaded over its area by the equivalent of atmos-pheric pressure-applied through the tie rods as illustrated·schemati-cally in Figure 11. Under the assumption that the pad acts as a simply supported-annular plate acted on by a uniform 14.7 psi presstn:'le load while restrained along its inner and outer circumferences, the maximum deflection equals 0.084 inches. The maximum stress induced by the loading is less than 1000 psi. The magnitude of this deflection is well within the limitations set by the criteria for the tie rod supporting structure, and indicates that the concrete pad has adequate stiffness under load. 4.7 Centre Post The requirements of beam injection and cyclotron physics restrict the permissible cross sectional area of the centre coltunn. The column functions as the inner wall for the vacuum tank and serves to connect the top and bottan. During the operating p~e of the cyclotron, the centre post is required to carry: (a) tne inner support reactions of all twelve primary radial members (b) the canponent·weight, carried by the pole tips of the six upper magnet cores (c) a portion of the magnet force acting between magnet halves (d) a part of the atmospheric pressure load These loads all induce canpression stress in the post. The total can-pressive load is found to be approximately 4200 kips. Studies indicate that a feasible design is possible when due considera-tion is given to: I I I. I-I I I I I I I I I I I I I I I (a) stress concentration and controlled yieldjng effects around the beam injection hole (b) a proper choice of the material to ensure adequate canpressive yield strength with non-magnetic behavior (c) the problems of manufacture involved for a member of this shape and size 12. Several materials to satisfy the above criteria have been investigated. The most pranising is Inconel 718 which in the age hardened condition exhibits the following properties: Ultimate Tensile Stren~h Yield Strength Elo~ation. 211,000 p. s. i. 174,000 p. s. i. 23% Magnetic Permeability 1.001 The estimated weight of the centre post in this material is 500 pounds with an estimated installed cost of $3,000. Further investigation during subsequent phases of the program may lead to cheaper alterna-tives. Also, at that time, it may be desirable to conduct laboratory tests on scale models of the centre post in order to verify analytical conclusions. The centre post loads are introduced through the upper magnet core tips in the initial design concept. The need for limited density of ferro-magnetic material around the centre post results in high stresses in the magnet pole tips when required to transmit the full centre load. A design utilizing a high strength structural spacer ring composed of very low permeability material is suggested. Its function would be to remove the centre loads from the magnet pole tips and transfer these loads to the centre post. It would also serve as a support for the magnet tips. 4.8 VacuUJn Tan1< Concepts 4.8.1 Pressure Effects on Tan1< The vacuum tank top and bottan has been analytically checked when acted upon by the full atmospheric pressure. A 7 I 8 inch thick plate supported by tie rods spaced not greater than 33 inches apart has a maximum de..., flection relative to support of 0.057 inches and is subject to a maxi-mum stress of 4,500 psi. The dilation effects in the plane of the plate are found to be negligible. The in-plane buckling effects on the plate can be neglected with risk of introducing only a 0.2 per cent error. It is also found that tie rod loading is quite insensitive to axial misalignment of the tie rod ends when supporting the evacuated vacuum tank. I I I. I-I I I I I I I I I I I I I I I 4.8.2 Thennal Effects on Tank A requirement to heat the vacuum tank to a temperature of 200°C results in uniform expansion of the vacuum tank top and bottan, and also introduces thermal gradients in the tank surface due to heat transfer effects induced by the tie rods. The unifonn expansion can be accanodated by providing an adequate clearance around the tie rods where they pass through the magnets. About 1/ 2 inch will be necessary. To avoid sealing problems .. due. to relative motions between vacuum tank body and cover, provision for adequately heating the outer circumferential wall of the tank is advisable. Temperature gradients between tank cover and wall are then avoided. The thermal gradients in the upper and lower surface of the tank 13. are determined assuming that the heating coils form an annular ring around each tie rod connection and that the tank acts as a thermally conducting body while the tie rod serves as a conductive, convective and radiative heat sink. These thermal gradients when found are analyzed to determine the stress they induce in the vacuum tank sur-face. Calculations show a thennal gradient of less than 40°C in the vacuum tank. This occurs on a disk-shaped plate around each tie rod connec-tion and is measured between the circumference and the centre of the disk. The maximum thermal stresses induced by this gradient are less than 10,000 psi radially and circumferentially. It can be concluded that heating to 200°C should not ove~stress the vacuum tank. As a result of these considerations and the pressure effects, it is clear that 7/8 inch plate supported by tie rods spaced at not greater than 33 inches is adequate for the vacuum tank structure. 4.9 Protective Devices The necessity of protecting the coils mounted on the top and bottan of the vacuum tank fran being crushed when the tank is vented or when the upper structure is elevated, led to the concept of providing mech-anical stops to limit the movement of the vacuum tank relative to the faces of the magnets under these conditions. Further advantages be-came apparent during development of this concept as follows: (a) Mechanical stops will be invaluable during erection and carmissioning of the cyclotron. During the period when the vacuum tank lid and bottan are to be adjusted to their horizontal plane under vacuum loading, it will be difficult, if not impossible, to adjust the tie rods correctly the first time. It will be reassuring to have coil protection when this adjustment is taking place. The mechanical stops under these circumstances offer very cheap insurance against costly coil damage. I I I. I-I I I I I I_ I I I I I I I I I (b) It should be possible to allow greater deflection in the primary structure with attendant decrease in beam depth, resulting in savings in building costs by reducing building height. 4.9.1 Selection of Suitable Stops The following methods of protective devices were considered: (a) (b) (c) Stainless steel blocks welded to the tank adjacent to the secondary coils and just slightly higher than the coil diameter. These would be spaced strategically throughout the coil system. Stainless steel blocks as in (a) above except these would bolt to the tank and would also act as coil attachnent blocks. "flange stops" used in conjunction with tie rod attachment points. This system is shown in Drawing 60001. 14. For the three concepts considered, the "flange stop" system (c) ap-pears to be the most appropriate and practical. It is chosen because there will be a need for same type of tie r:xi attachment footing to the vacuum tank, regardless of which protective system is chosen. It is simple to dimension such a footing so that it will function also as a stop. This approach saves the expense of the separate block system in (a) and (b) above. However, a few stainless steel blocks may be needed for coil protection near and under the edge of the magnet and the edge of the tank lid where the "flange stops" are ineffective. The reason for this is that during an elevating condition, the tank lid surface assumes the elastic curvature of the supporting structure, while the magr.et face remains aJmost plane. This tends to deflect the outer edges of the tank upward relative to the magnet and could cause the coils to impinge on the magnets around the tank periphery, with consequent coil damage. Also, the blocks may be required near the mag-net edge to protect against differential springback between rods in the valleys and rods constrained by "flange stops" under the hills. The rods in the valleys will return to zero deflection when the tank • is vented, but the rods passing through the magnet may not, resulting in a portion of tank surface in the valleys being elevated above that portion under the hills. This effect may cause the coils to be pushed into the magnet edge. The operation of the "flange stops" is simple. A recess is machined into the magnet where a tie rod passes through. The footing of the tie rod loosely fits within the recess. The shoulder depths of the recess and the footing are dimensioned such that when the tank is ventE1(i and the tie rods contract, the footing shoulder will rest against the recess shoulder before the coils touch the magnet faces. I I I I I I I_ I I I I I I I I I 4.9.2 Analysis of Sn:_U:c::tu~ for Increased Deflection·· To integrate the stops with the support structure implies a relax-ing of the structural deflection criteria. Without stops, the overall structural deflection must be less than 1/4 inch minus manufacturing tolerance. With stops, however, greater allowable structural deflections may be tolerated, suggesting a shallower beam design. A "worst case" type analysis was made in order to study the effect of incorporating the "flange stops" into the vacuum tank support systan. Three primary structure maximum deflections were chosen: 0.050 inches, 0.100 inches and 0.150 inches. The same secondary structures and tie rods are used for all three cases and therefore their deflections remain constant. Three "worst case" assumptions were made in the analysis. The first is that the magnet face takes on a similar elastic curve to the primary beam and has a maximum deflection of 0.100 inches. Lack of infonnation at the time dictated that this assumption be made. Not to include it would suggest that the magnet faces are contour machined to lie perfectly flat in the horizontal plane. At this stage, this suggestion does not appear reasonable. The second assumption is the inclusion of 0.063 inches manufacturing tolerance in the analysis. This dimension may.be considerably less than assumed, or indeed may be entirely adjusted out, but to neglect it would be unrealistic. In order to eliminate this value in the calculation, it would be necessary to assume that each and every tie rod footing base lies on the same horizontal plane after adjusiJ'llent, which is improbable. The third assumption is that the primary and secondary structure returns 100% to its unloaded position when the tank is vented, leaving the tie rods to accept all of the springback deflections. This assumption appears realistic for the 0.050 inches maximum deflection case because the primary structure required is very stiff. However, for the 0.150 inches deflection case with the increased structural flexibility, the assumption is conservative. For this case, the structure will springba.ck to sane residually deflected position in balance with any residually loaded tie rods when the tank is vented and the stops are in use. In order to 15. allow for canparison between the three cases, this asstunption is maintained throughout the analysis, keel)ing in mind that it becanes less valid as greater primary structural flexibility is encountered. A canplete analysis based on allowing residual deflection of the primary structure is canplex and not within the scope of this study. In general, the·results of the analysis indicated that the system can be affected in three ways depending on the degree of structural deflection, as follows: I I I~ I-I I I I I I_ I I I I I I I I I 16. 1. The tie :rods are equal length when the tank is under vacuum . conditions. When the tank is vented, a certain number return to zero stress and deflection while the remainder are pre-vented fran doing so by the "flange stops". The tie :rods returning to zero stress and deflection are those which pass. through the valleys, plus those which pass through the ·hills but have enough clearance to return 100% to the unloaded position. The remaining tie :rods are prevented fran return-ing to zero condition by virtue of the stops caning into play. These particular :rods attempt to prevent the primary and secondary structure fran returning to the unloaded position. This causes the load per :rod to increase as there are fewer rods attempting to restrain the structure. Thus, the resi-dually loaded tie :rods can have higher stresses in the non-operating condition than in the operating one. This effect becanes more pronounced as greater structural deflections are allowed. 2. The sum of the residual stresses in the tie :rods over their total area is equal to a resultant.upward force on the upper magnet sectors. If this force exceeds the weight of the sectors, then the magnet may be dislodged as a result. 3. A differential deflection will result near the magnet edges due to tie :rods in the valleys returning to zero deflection and the :rods in the hills being prevented fran doing so by the stops. The results of this analysis for three assumed deflections are summarized for each case below. The numbers represent approximate values only, and are based on the "worst case" analysis described above. Maximum Primary Structure Deflection . 050 in. . 100 in . .150 in . Number of tie :rods with residual stress 6 12 19 (1-1/2 diameter assumed) Maximum stress (psi) 10,000 23,000 42,000 (in 2 rods) (in 3 rods) (in 1 rod) Total upward force per magnet sector (kips) Negligible 290 680 (WI' of Sector= 516 kips) Possible to dislodge magnet No No Yes Hill to valley deflection Negligible 0.060 0.120 differential (max. - inches) Ccmnent A B C I I I-I-I I I I I I_ I I I I I I I I I 5. A. "Flange stops" are recamrended even for this small deflection due to manufacturing variations. B. The ratio of magnet weight to upward force is 1. 7 5 , which prevents uplift of magnet. A higher strength rod may be needed in some areas to contend with high stresses. C. The total upward force per sector is greater· than·· the -weight of the sector implying that the magnet would be ·dislodged~.-In addition, sane of the tie rod stresses are higher than, ortlinary structural steels could safely accarmodate. • • Both of these facts suggest that 0.150 inches primary structural deflection is too great. However, the assumption that the • primary structure springs back 100% to its unloaded position is not -really valid for this case. The actual stresses and resulting upward force would be less than detennined by this study. Further detailed analysis taki."'lg primary structural flexibility into account would be necessary to place a limit to the allc,r..,Table·deflection. MANUFACTURING CONSIDERATIONS· 5.1 Vacuum Tank 17. The lid and bottan of the vacuum tank may be constructed fran 7/8 inch type 304 stainless steel plate. Because of the limited amount of space available between the magnet gap, the manufacturing varia-tions becane important and must ·be reduced to the practical minimum. Two effects were considered; first, the plate waviness and thickness. tolerance of rolled plate as supplied fran steel mill and second, the angular and step misalignment due to welding fabrication of butted flat plates; • 5.1.1 Fabrication Tolerances Angular and step misalignment of butt welded stainless steel plate can be reduced to a minimum by special welding techniques. The maxi-mum step misalignment can be reduced to the difference of thiclmess between the two plates. For 7 / 8 inch plate, the value would be a maximum of O. 074 inches. Angular misalignments can be reduced to the extent that they are no longer significant with respect to other tol-erance values. To accanplish this, however, a special manual welding technique must be used in conjunction with structural hold-downs called "strong-backs". This technique entails the use of small diameter weld rod desposited in a pattern designed to minimize thermal distortion. 5.1.2 Mill Tolerances The regular flatness tolerance on 7/8 inch plate, 114 inches by 27 feet, is 15/16 inches out of flat on 15 feet of plate. Special flatness of 1/8 inch maximum deviation out of flat throughout the plate can be neg9tiated with steel suppliers. The anticipated extra for this is approximately $1. 75 per 100 lb. of material. Present stainless steel plat~ mill prices run about $80.00 per 100 lb. It appears econanically feasible to negotiate a special tolerance since the increase in cost is only about 2% for an excellent increase in flatness. I I I-I-I I I I I I_ I I I I I I I I I 6. The regular thickness tolerance is+ 0.064 inches - 0.010 inches. It is not known if special thickness tolerances can be negotiated.. 5.1.3 Overall Manufacturing Tolerances 18. Since the top and oottan of the vacutun tank will be made plane and parallel at the tie rod connection points when the system·is·ad-justed under vacutun conditions, the deviation of the-plate~fran-.• • flatness is of ·concern only between these attachnent ·points; - -'fh~--1/8 inch departure fran flatness is specified as a maximtun ·for ·the .. • whole plate. It is unlikely that this maximtun deviation 'WOUld occur between two adjacent tie rods, ·and •it is even more unlikely that it would occur in conjunction with the maximtun step misalignment of 0.074 inches due to varying plate thicknesses and manufacturing discrepancies. Therefore, a total departure fran theoretical flatness of 1/8 inch between any two tie rods is accepted as being statistically reasonable. 5.2 Support Structure The support structure is a structural steel fabricated straight beam arrangement. The primary structure consists of a centre ring, fran which 12 fabricated wide flange beams extend radially. The beams are welded to the centre ring to make a hanogeneous structure weighing approximately 100 tons. The secondary structure consist~ of rolled channels for tie rod support and rolled (or possibly weld fabricated) wide flange beams for magnet support. The secondary structure is bolted to the primary so that the largest single lift for the cranes will not exceed approximately 100 tons. There is no major difficulty in fabrication or erection of the support structure. In addition, requirements are well within the capabilities of the larger local manufacturing finns. INVESTIGATION OF JACKS 6.1 General In both the leveling and elevating jack systems, the total load or load per jack is rather high. This eliminates most of the carmercially available mechanical or hydraulic jacks suitable for lower load cap-acities. HCMeyer, there are sane canmercial jacks available as stand..,. ard units or specially designed which would be in.range of the required load capacity. These and the specially designed devices are described as follows. 6.1.1 Ball Screw Jacks Ball screw worm gear jacks are manufactured up to 50 tons capacity by the Canadian Duff Norton Canpany and Joyce Cridland of Dayton, Ohio. They can be designed and manufactured over the 5 0 ton capacity, but the price per ton of lifting capacity increases con-.. siderably and delivery time required may be a problem. I I I-I-I I I I I I I I I I I I 19. Because of the highly efficient ball screw assembly, overall effici-ency of a 50 ton jack is in the range of 45% to 68%, depending on over-all jack ratio. At this efficiency range, the jacks themselves are not self locking although when canbined with wonn gear speed reducers, the overall system will be. However, a brake is recanmended ·for holding and positioning the loads._ . The basic load capacity can _be :i,ncre.ased to some extent where the number of operating cycles of the jacks is comparatively small as for the leveling and elevating systems. How-ever, this would reduce the basic factor of safety in strength and thus reduce reliability of the units. 6.1.2 Acne Screw Jacks The Acme screw worm gear jacks are similar in construction to the ball screw jacks and are manufactured by the same two companies up to 250 tons capacity. The efficiency of these jacks is much lower than the ball screw assembly. For a 150 ton jack the overall jack efficiency is in the range of 9% to 14% depending on overall jack ratio. The overload capacity can be increased considerably over the standard rating by replacing the anti-friction thrust bearings with plain bear-ings and using higher strength materials. However, replacement of anti-friction thrust bearings by plain bearings reduces the efficiency still further. In this case they are suitable only for holding static loads rather than raising or lowering. Al though the jacks are self locking, addition of brakes would ensure better stopping accuracy. 6.1.3 Motor Driven Wedges Specially designed motor driven wedges were considered in canbination with some other lifting devices for magnet leveling systems. These could be used for supporting the load when the leveling adjustment has been accomplished. The motorized wedge mechanism (see Drawing 60001) consists of a guide frame with the upper wedge guided vertically and the lower wedge guided horizontally. It is driven by a screw coupled to a small gear motor. With a wedge angle of approximately 10 degrees, an accurately adjustable stop for fine adjustment could be·obtained econanically in comparison with the screw jacks. 6.1.4 Hydraulic Actuators One of the most canmonly used devices for short stroke raising or lowering of heavy loads are hydraulic actuators. There are many canmercially available types, but for the sizes required in this case, the number of suppliers becomes limited. For 2000 and 3000 psi working pressures, actuators are available up to 12 or 14 inches diameter. Larger sizes can be obtained to special designs but the price per ton capacity then increases considerably. I I I-1-1 I I I I I_ I I I I I I I_ I I 20. 6.2 Alternative Jack Arrangements Several alternative arrangements were considered for magnet leveling and elevating jack systems. Most of these arrangements would satisfy the basic perfonnance requirements but considering reliability in case of overloaded jacks or unreasonably high cost of specially designed high capacity units, only a few were selected for final evaluation. In the magnet leveling system, the load per jack is approximately 258 tons. Using a specially designed ball screw jack of that capacity, the price per jack would be unreasonably high. Even canbining two 150 ton ball screw jacks in pairs to raise that load would cost $7,000 per jack or $178,000 for the canplete jack system with drives, but without controls. Even if 250 ton Acme screw jacks were used at a naninal cost of $8,600 each the total cost of the system with drives would be $122,000 without controls. The high cost of the above ar-rangements does not make them very attractive, and therefore these are not considered for further evaluation. Two 150 ton Acme screw jacks canbined in pairs would offer a reason-able arrangement at a total cost of $56,000 including drives but without controls. However, this cost does not include a load equal-izing member in the form of a beam which is required for load sharing between the two jacks. A similar arrangement, but utilizing a lever instead of a beam is considered, This would satisfy the same require-ments, but uses only one 150 ton Acme screw jack per unit. Since this is irore econanical than the double jack arrangement, it is selected for final evaluation and the double jack arrangement is not further considered. The load rating for the elevating jacks is 129 tons per jack based on design criteria, but during the course of study, this has increased to about 145 tons per jack. Here again, cost of ball screw jacks for that rating is prohibitive. On the other hand, canbining two 5 0 ton ball screw jacks to raise that load would be econanical at an esti-mated approximate total cost of $49,000 for the jacking system with drives without controls, but the overload is much too high where a reliable system is essential. The same applies to the single 100 ton or double 50 ton Acme screw jack arrangement with estimated costs of $62,000 and $50,000 respectively for each canplete arrangement with drives. For this reason the above arrangements are not further con-sidered. While hydraulic actuators are well suited for raising or looering high loads, they are rather poor for precise position holding because of piston leakage unless servoed. Normally with cast iron piston rings which require minimum maintenance, there is always some leakage, If some elastomer piston seals are used, leakage can be eliminated, but regular replacement of seals is required to maintain perfonnance, However, hydraulic actuators were considered in combination with Acme screw jacks for the elevating systems. The load could be raised and lowered with the hydraulic actuators and at the top of the stroke held by the screw jacks. Cost of such a system using 12 inch diameter ac-tuators with modified 50 ton or 100 ton Acme screw jacks would range fran $55,000 to $62,000 without the hydraulic power supplies and cont-rols. Since a rather canplicated servo control system would be I I I-1-1 I I I I I_ I I I I I I I I I 21. required to synchronize the strokes of all the actuators during raising and lowering, total cost of the system is rather excessive. After disregarding most of the jack arrangements as unsuitable or_ too costly, the following few have been evaluated for reccmnend-ation. 6. 2 .1 Magnet Leveling Jack Systems __ (a) (b) 24 - 12 inch diameter hydraulic actuators canbined.in pairs at each load point and used in conjunction with motor driven wedge mechanisms. The hydraulic actuators would be used only for raising and lc::Mering the load and the wedge mechanism for holding the load of the magnet cores in the adjusted position. To raise a 258 ton load with 2 - 12 inch diameter actuators, a working pressure of 2300 psi is .required. • For the required· speed, of_ 0.4 inches ·per minute, a flow rate of 0.40 ·gpn is. required. To overcane friction losses in the hydraulic cylinders plus line pressure losses and to canpensate for piston leakage, a 3000 psi hydraulic system with approximately 1 gpn delivery would be required for safe operation. 12 - 150 ton Acme screw jacks driven individually by 10 _hp motors@ 1750 rpn through speed reducers and canbined with lever mechanisms (see Drawing 60001 version C) would produce a vertical velocity at load point of approximately O. 44 inches per minute. The drive power could be reduced if a lower speed is acceptable. 6.2.2 Elevating Jack System Only .one jack arrangement was selected for final evaluation. This arrangement canprises 12 - 150 ton Acne screw jacks where each jack is individually driven by a naninal 15 hp motor at 1750 rpm through a gear reducer to obtain an elevating speed of 2.1 inches per minute. The motor would be equipped with a brake to obtain a high consis-tency in stopping accuracy. 6.3 Control Systems Several alternative control systems for the magnet core leveling and elevation systems have been considered. Since most of the mechanical jacking system alternative arrangements are identical fran the con-trols point of view, alternative control systems were studied inde-pendently fran the jacking systems. Only the hydranechanical jacking systems require different considerations fran the control point of view. I I-I· I I I I I 1-1 I I I I I .. I_ I I 6.3.1 Control Systems for Leveling Jacks The basic requirement of this control system is to drive and ac-curately control each jack individually. Since only one jack 22. would be operated at any one time, only one canmon control system is required with provisions for switching to any one of the 12 individual jack drives. Toe naninal stroke of 2 inches for instal-lation adjustment does not have to be closely controlled but the fine adjustment within 0.25 inches range during operation of the cyclotron must be controlled within plus or minus 0.010 inches accuracy. Two alternative systems have been considered: one for a hydro-mechanical jacking system where hydraulic actuators are used for raising and lowering the load and the motor driven wedge is used as a reference stop for holding the dead weight load and the other for a mechanical Acme screw jacking system. (a) (b) Toe control system for the hydro-mechanical system is more canplicated because of the two system a.I'I'ange-ment. This control system as shown in Figure 12 is based on conventional servo position relay systems. Each jack unit has its own separate control loop. After adjustment of the sensing transducers of each system, the references are so adjusted that the hydraulic system leads the wedge stop by a predetermined distance (say 0,010 to 0.020 inches) in terms of elevation, This allows the wedge to be free of any load during the adjustment operation. Toe control system for the mechanical screw jack arrangement is a conventional servo position relay system as shown in Figure 13. Once the coarse adjustment of the magnet cores has been canpleted, the transducers of each jack are centred within their span and reference synchro and position dial is also adjusted accordingly. 6.3.2 Control Systems for Elevating Jacks Six possible approaches to the problem of driving 12 jacks simul-taneously to an acceptable tolerance were investigated. Fran structural considerations this tolerance was established as plus or minus 0.025 inches frc:m the nominal position. Each control system was considered to be a form of position control, either manual, open loop, or semi-autanatic closed loop. Considering each type of system, there is a problem of ensuring that the error between any given number of jacks does not exceed the established limit. In the case of the open loop systems that are electrically locked there is ambiguity at the high speed end of the drive, i.e. at the motor. Initial alignment would ensure that the rotors of the motors are all in phase and locked within 90 degrees electrically of each other, but power inter-ruption or friction effects during stopping of the drives over a I I I-I I I I I I I I I I I I I 23. period of time after several operations could cause serious.mis~ alignment between the individual jacks and this would not be acceptable. The various control systems considered are as follows: (a) Partial Selsyn System (b) (c) It is possible to run several wound rotor motors together . so that they will operate in phase alignment. Primaries are connected together and secondaries to a camion exter-nal resistance. This arrangement has sane difficulties in that the rotor resistance needs to be adjusted to en-sure at least 10% slip in driving up. Sufficient load is required to keep the rotor retarded. The worst situation will occur when the jacks are lower-ing the load which would reduce loading of the motors. There is a risk that any one drive can be so lightly loaded that its motor will speed up and get out of step. Although econanical, this arrangement is not considered safe. Standard Selsyn System Full power selsyn system as shown in Figure 14 is considered feasible as considerable control can be achieved to ensure the necessary rotor phase pull out torque. In this case, the stator excitation phase rotation is opposite to that of the rotor rotation. Each jack drive rotor will have at least 200% to 300% phase lock torque depending whether the receiving rotor is ahead or behind the transmitting rotor. Fran a performance point of view, this system is ac-ceptable, but the estimate cost of controls with motors is approximately $42,000. Synchronous Reluctance Drive This system has recently been used for dry docking facilities equipped with multi-drive (about 20) lifts of 15 hp each. Squirrel cage motors with slots cut in the rotors fonning a salient pole type con-figuration are used. Synchronous speed is reached in a few cycles. l.Dcked rotor and pull out torque is in the order of 150%. Inrush currents are approx-imately 600%. Sane difficulties have been experienced because of the regulation of the power source. Power factor correction and time delay sequence starting has been necessary. No detailed data are available, but fran the performance history this system is not con-sidered suitable. I I I I I I I I I I I I I I I I I (d) (e) (f) 24. Closed Loop Control System This system as shown in Figure 15 utilizes standard gearmotor drives for each jack. Control of monitoring equipnent keeps the drives in step with each other. A camnon moving reference is used to provide the track-ing guide. This system autanatically keeps all jacks in step and prevents misaligrment. If one or more jacks lag or lead the others, the system autanatically.stops and re-starts until all jacks are again in synchronism. For a jacking system with an output speed of 2 inches per minute driven by approximately 10 h.p. induction motors equipped with electranagnetic disc brakes .of _so ft. lb. capacity, stopping of the system can be acheived in approximately 150 radians at the high speed end or 0.027 inches of jack travel. This can be further re-duced if necessary. Fran a performance point of view, this system would satisfy the requi.renents at a cost of approximately $22,000 including the squirrel cage motors. Line Shaft Drives - Two Jacks Linked per Motor This system is basically the same as described above in (d) except that jacks are linked in pairs with shaft-ing and driven by single motors. This would reduce the number of drives fran 12 to 6. This concept does not appear attractive because of space considerations for line shafting and drive. The estimated cost of $21,000. including motors, but without installation, is quite reasonable. However, cost of shafting, couplings, and drive support base would increase this considerably. Line Shaft Drives - Three Jacks Linked per Motor-Again, basically the control system is as in (d) and (e) but the number of drive motors has been reduced to four. Al though the cost of motors and controls amount only to $16,000, the additional cost of line shafts and couplings, plus difficulties to accanodate shafting without interference with other equipnent rules out this concept for further consideration. 6.4 Operating Envirorment Essentially, both types of screw jacks and the hydraulic actuators are suitable for taking vertical loads only in both the elevating and leveling systems environments. Extensive side loads very likely would jam or even damage the jacks to an extent where operation of the system would be interrupted. I I 1-1· I I I I I I I_ I I I I I I I I 25. Examining the elevating system, it is evident that temperature .changes of the structure when supported on the jacks in the elevated condition would cause relative movements resulting in side loads on the jacks. In the case of the levelling system, temperature variation of the mag-net cores would result in radial movements which would also impose high side loads on the jacks. Also, earthquake conditions would subject the jacks to side loads unless provisions have been made to eHminate effects of these conditions. The extent of side loads has not been investigated in this study.nor provisions included to protect the jacking systems against such loads. It will be necessary to investigate these conditions and make provisions for protection of both jacking systems. 6.5 Evaluation of Alternative Jacking Systems 6.5.1 General Fran the foregoing discussions of individual types of jacks and canparison of alternative jack arrangements, two alternative jack arrangements for the magnet leveling system and one mechanical ar-rangement with two control options for elevating jack system was selected for final evaluation. In making the first selection of the alternative arrangement for recanmended system, the following aspects were considered: (a) perfonnance and reliability of system (b) ease of operation and maintenance (c) cost While cost is a very important factor and minimum maintenance is very desirable, the most important factor for both the leveling and the elevating jack systems is reliability. For the environment conditions, where extremely heavy loads must be supported and ele-vated by remote operation, and where accessibility to the canponents of these systems is severely restricted because of radiation hazards, ... it is absolutely essential that these systems are fail safe and re-liable in operation. 6.5.2 Magnet Leveling Jack System The hydro-mechanical system, utilizing hydraulic actuators and motor driven wedge stops would cost approximately $9,000 more than the 150 ton Acme screw jack arrangement. Although the first system is more canplicated and has a more canplex control system than the other, the fact that the load raising mechanisms (hydraulic actuators) are subjected to the loads only during the operation and the wedge. stops are relieved fran the load during their adjustment make this system more reliable than the other alternative. Even if more main-tenance may be required on the hydraulic actuators in periodic rod I I I-I· I I I I I I_ I I I I I I I I I 7. 26. seal replacement, the actuators can be removed without any disturbance to the magnet leveling which is not the case with the 150 ton Acme screw jacks since they support the load continuously. Al though the simplicity of the 150 ton Acme screw jack system makes it very attrac-tive, the hydro-mechanical system is selected as the first choice for the above reasons. 6.5.3 Elevating Jack System Of all the considered alternative arrangements only one mechanical arrangement was selected as suitable fran reliability and cost point of view. Al though a ball screw jack arrangement appears very attrac-tive fran the higher efficiency point of view, the cost of adequate capacity jacks for this application is excessively high. The same mechanical arrangement was considered with two alternative control concepts and the cost of both systems canpared. Fran this canparison it is evident that the arrangement with closed loop control system is more econanical by an amount of $12,000. From a perfor-mance point of view, both control systems would satisfy the require-ments. This indicates that the preference should be given to the closed loop control system. COST ESTIMATES 7.1 Supporting Structures The estmia.ted installed cost of the straight beam concept is can-pared to the optimum an:h structure as follows: (a) Straight Beam Beam depth (b) Total structural weight Cost Arch Operating arch rise Elevating an:h rise Depth operating an:h section Depth elevating arch section Total structural weight Cost Net saving (straight beam vs arch) 60 ins. or 90 ins. 297,000 lbs. $84,500 105 ins. 200 ins. 30 ins. 18 ins. 346,000 lbs. $105,000 $20,500 I I I-1-1 I I I I I_ I I I I I I I I I 27. N.B. kn estimated cost of $3000 for the Inconel 718 centre post has been included in the above figures. In addition to the above savings in the installed cost of the straight beam canpared to the arch concept, further econanies are effected by reduction of building height as follows: (a) Straight Beam - 60 inches deep (b) Height of arch structure Height of straight beam 209 ins. 60 ins. Height saving 149 ins. Building cost saving@ $8,500 per vertical foot (data fran TRIUMF) $105,500 Add manufacturing saving 20,500 Total saving $126,000 Straight Beam - 90 inches deep Height saving Building cost saving@ $8,500 per vertical foot (data fran 119 inches TRIUMF) $84,000 Add manufacturing saving 20,500 Total saving $104,500 N.B. The deflection of 0.10 inches for the 60 inch beam ccmpared to 0.050 inches for the 90 inch beam can probably be tolerated when used in conjunction with the "flange stops" discussed in Section 4,9. 7.2 Jacks and Controls 7.2.1 General For the magnet leveling and elevating jack systems and their controls, cost estimates have been prepared for the equipnent only and no allow-ance has been made for installation and tests. Since the scope of this study was restricted to certain areas only of the cyclotron, realistic cost estimates could not be established for the provisions I I I-I· I I I I I_ I I I I I I I I I 28. required to incorporate the jack systen into the cyclotron equipnent. These provisions, such as support bases or foUMations for jacks and their attachment to the equi:i;:rnent, would be included in the cost· of. the equi:i;:rnent affected by the jack systens. Therefore, a true assess.,. ment would not be made if only the cost of installation was added .to the cost of equipnent. However, the cost of assembling control .. equip-. ment, instrumentation and hydraulic equipnent has been included. The following cost estimates cover only those arrangements which have been selected for final evaluation or reccmnendations. 7.2.2 Magnet Leveling Jack Systans (a) Hydro-mechanical jack arrangement 24 - 12 inch diameter hydraulic actuators $31,000 12 - wedge mechanisms 18,000 12 - geared motor drives with brakes 5,000 Control systen including 3000 psi hydraulic power supply, hydraulic tubing and assembling <»f equip-ment • 19,000 (b) 150 ton Aane screw jack arrangenent with levers 12 - 150 ton Aane screw jacks 12 - speed reducers 12 - 10 h.p. motors with brakes Lever and support bracket a.ITa.ngement Controls and instrumentation 7.2.3 Elevating Jack Systens (a) 150 ton Acrne screw jack ar.rangenent with closed loop control systen 12 - 150 ton Aane screw jacks 12 - speed reducers $73,000 $19,000 13,500 5,000 15,000 11,500 $64,000 $43,500 22,000 I I I-1-1 I I I I I_ I I I I I I I I I 8. 29. 12 - 13 h. p. 100tors with brakes $ 7,500 Closed loop control system 23,000 $96,000 (b) 150 ton Acme screw jack system with standard selsyn control system 12 - 150 ton Acme screw jacks $43,500 12 - speed reducers 22,000 Selsyn control system with drive motors and brakes 42,500 Total $108,000 All cost estimates are based on prices in effect at date of this report and do not include any Federal or Provincial taxes nor provision for e:scalation. RECOMMENDATIONS 8.1 Supporting Structures A straight beam concept embodying 12 - 90 inch deep fabricated I-beams spaced radially and welded to a stiff centre ring is recamnended for the primary structure to give a total weight of approximately 100 tons. Secondary structure consisting of channel sections to support the tie rods and I-beam sections to provide lifting structure for the upper poles of the magnet will be bolted to the primary structure. Tie rods of 1-1/2 inches diameter, spaced not greater than 33 inches will re-strain the vacuum tank during operation. To withstand the high stresses in the centre post, Inconel 718 is recamnended for this member. Also, consideration should be given to a high strength, low penneability spacer ring at the centre of the structure to remove the loads fran the magnet pole tips and transfer these through the centre post into the foundations. The use of protective devices or "flange stops" to prevent crushing of the secondary coils is suggested. A further detailed study of supporting structure deflections when used in conjunction with the stops is recarmended in order that beam depth can be optimized to effect econanies in beam manufacture and building costs due to de-creased beam depth. I I I I I I I 1. I I I I I I I_ I I 30. 8.2 Jack Systems 8.2.1 Magnet Leveling Jack System It is recamnended that the hydro-mechanical jack arrangement with its control system as described below be adopted as a basis for detail design to be incorporated in the cyclotron system, as follows: Twelve jack units each consisting of two, 12 inch diameter hydraulic actuators suitable for 3000 psi operating pressure canbined with one motor drive wedge assembly as shown on Drawing No. 60001 (Version C). The wedge mechanism to be operated with 1/4 h.p. geared motor, Boston MW126-300 or similar with disc brake, coupled to a 1-1/2 inch diameter, 3/8 pitch Acme screw to operate a wedge nechanism with approximately 10 degrees wedge angle. One comnon hydraulic power unit suitable for delivery of approximately 1 g.p.m. at 3000 psi with necessary valves and line connections to all the hydraulic actuators. A closed loop control system generally as shown in Figure 12. It is recaranended that detailed studies be made of the horizontal movements of the magnet cores due to temperature effect or magnet forces and the resulting side load on the wedge mechanism be considered in final design. If necessary, suitable provision should be made for protection of the magnet leveling system fran these and other loads such as seismic. Further consideration may be given to the alter-native jack arrangement using 150 ton A<::IC"e. screw jacks if this system should offer advantages over the recamnended system when the side load effects have been evaluated. 8.2.2 Elevating Jack System It is recamnended that the jack arrangement with its control system as described below be adopted, as follows: Twelve, 150 ton Acme screw jacks driven individually by standard 3 phase, 60 cycle motors rated at 15 h.p. @ 17000 r.p.m. through wonn gear speed reducers. The motors to have locked rotor torque rating of 200% to 300% of normal full load torque rating to overcane the static friction in the jack screws and gear train. The motors to be fitted with disc brakes for accurate control and stopping distance. The jack system to be operated and controlled for synchronous op-eration with a closed loop control system as shown in Figure 14. Further studies should be made of the effect of possible lateral forces acting on the jacks when they are supporting the vertical load. Sui table provisions for eliminating lateral movements, par-- ticularly those due to seismic loads, should be made. ~I I 1. 1-1 I I I I I. I I I I I I I I I SYSTEM (a) COMPLETE SYSTEM ELEVATING OPERATING Fl G.1 a I I I-I-I I I I I .. I_ I I I I I I I I I SYSTEM (b) COMPLETE SYSTEM ELEVATING OPERATING FIG. I b I I I-1-1 I I I I . I_ .I I I I I ,, I I I SYSTEM (C) COMPLETE SYSTEM ELEVATING -- -~-----·-·-·----····· OPERATING FIG.I C I I I-1-1 I I I I --I_ I I I I I I I I I MAXIMUM DEFLECTION OF ARCH VERSUS ARCH RISE (,~]t 0·3-----w (/) a: SPAN • 40 IN'l O·I f-----+------~-_..;;::,,.,--=:-------+- ~RI NG -~-so,J AREA ---'-120 IN2 O 60 90 120 ARCH RISE (IN) FIG. 2 I I MAXIMUM DEFLECTION OF ARCH VEASUS 1-RESTRAINING RING AREA 1-(,~J I I 0·30 I 0·25 I I 0·20 I_ 60tN I 0·15 I ARCH RISE 0·10 90\N I 120,w I 0•05 I ··I O 40 80 120 I RING AREA (INJ I FIG. 3 I 1t! TRANSLATION OF OUTER-ARCH SUPPORT VERSUS RESTRAINING RING AREA 0·15Qi----;----------t----------+----------t 601N 120 IN 0·025----------------4------ --O 40 80 120 RING AREA (1N1) FlG.4 (l~r TRANSLATION OF OUTER-ARCH SUPPORT VERSUS ARCH RISE 0·150t--t------,-.--ll"c----+---------+------ 4O1N2 O·IOOt----,,.f---------+------,----+---~----1 RING AREA a OD50t-t-----_;::~~------+--8_0_1N--+---~ 0·0251--t-------+-------+------...J O 60 90 120 ARCH RISE (IN) FIG.5 .. /41 1/ I I I I u.J a: I i :::::, i I-( I I I () • cc ; :::::, \ \ ,-.... a:: r,f) I-\ \ l/) \ J: () a: ~\ <{ <!) • • LL (!) 0 0 (X) -~ -,-.... z LL 0 -t-fi// w I I ' _J w I I I I u -t-<( & ( f = Is • w ~\ \~ I~ co I ,-.... u r<) (/) \ \ \ ~\ ' • IO 0 • 0 0 N I I I~ 1-1 I I I I I. I I I I I I I_ I I MOMENT OF INERTIA VALUES VERSUS DEFLECTION OF STRAIGHT BEAM OF 95 % CONFIDENCE BAND '\ \ / , 100,000-----~, ---------+--------! \ / \~- )_ \ so, ooo..,__----+-----4-+----~:.+-,.L.-....,..C,......-+------t / . _/ -· "-.-/ _/ --· '<. / / . • _/ / . ~/ -· /, , / / / o.os" 0.1011 o.1s~ 0.20· o.2s· MAXIMUM BEAM DEFLECTION (IN) CASE I (IN4 ) A 5 (IN 2 ) AT (IN 2 ) L (IN) I 150,000 90 136 360 --2 1401000 6 7. 5 116 360 3 136,000 90 128 360 4 75,400 60 129 360 5 63,600 45 205 360 6 31,800 43 103 342 7 31,800 43 103 372 SHEAR AREA TOTAL AREA BEAM Lftll6TM FIG.7 - - - - - - - - - - ,- - -I I 1 ------' 212· X •·-- -- 23a• ' ·-- - I 204" --i 71.SK -- --, 'y 110· 63K -·-I 59.SK 135" -1 54K 100· - 44.5K 65" - I - I 33" I 34.4K 6. 23.4K _30" 12.85K -' t) 5.85K 10 11 I • 11' , . , , , , 1, Ir ', 't I,, I 2 3 4 5 6 7 8 9 ? -2 - 306" . --· -·· . ···--340 11 ---360 11 --•·--- - . --· --------·--•-----····------· SCHEMATIC OF STRAIGHT BEAM FIG. 8 I I I~ 1-1 I I I I I. I I I I ,1 ·1 I_ I I TOTAL STRUCTURAL WEIGHT VERSUS ARCH RISE (MAXIMUM DEFLECTION OF PRIMARY STRUCTURE• OS) 500 I 1 , ci~1t vsrRA1GHT sEAM -Wi~~T -(4s·o.~EP] • ..., .. -+----+----':~ 400----+------4---+- --------1--------- .... ---t-----¾ ARCH CONCEPT - TOTAL :;T~:--o..._ TURE WElG~T I -300---+---1---+----'-n-----1----1 ~----~--~-._J ..... 'o .... ARCH CONCEPT _PRIMARY STRUC - )-1 I TUF!E WEIGHT STRAIGHT BEAM -i;>ET~ J_ (60"DEEPJ 2001----;---+---+----+---+---~-~ 100----+------+---+-----+---+----+----I 20" 40" GO" BO" roo" 120" ARCH RISE FIG. 9 -------------------• 1 t I f 200" 105" 378" 376" 780" DOUBLE-ARCH CONCEPT -48• DEEP SECTION FIG. 10 I I I 1-1 I I I I I. I I I I I I .. I_ I I SCHEMATIC LOADING DIAGRAM OF LOWER CONCRETE SUPPORTING STRUCTURE SYMMETRICALLY ABOUT t_-D = 324 IN. DIA. d - 60 IN. DIA. t-3f>IN. 0 UNIFORM ATMOSPHERIC LOAD OF 14.7 P, S.I. R = OUTER CIRCUMFERENTIAL SUPPORT R1 - INNER CIRCUMFERENTIAL SUPPORT FIG. 11 -------- - - - - - - - - - -. ., REFERENCE SYMGMRO DETECTOR LOGIC . • I POWER IOTURN POSITION DIAL T AMP fOR I I I I I I I REFRENCE LVDT 1 DEMODULATOR REV F'OR-REV STARTER OTHER SOLENOID VALVESO AMP 1 DEMODULATOR/ HYDRAULIC L..!...J POWER SUPPLY 0 OTHER JACK DRIVES 0 GEAR MOTOR 0 SENSE SYNCHRO c 4WAY I MYDRAULIC SOLENOID I VALVE CLOSED LOOP CONTROL SYSTEM FOR LEVELING JACKS FIG.12 • MECHANICAL WEDGE -7 7 SENSE LVDT -- - .-,- - - - - - ,-~- - - - - - - -10 TURN REFERENCE SYNCHRO POSITION .Ir AMP DIAL DETECTOR LOGIC DEMODULATOR I OR 3 ct> POWER T T FOR REV FOR-REV STARTER 0 0 0 OTHER JACK DRIVES GEAR MOTOR SENSE SYNCHRO I I I II LIFT JACK I J CLOSED LOOP CONTROL SYSTEM FOR LEVELING JACKS FIG. 13 • .STANDARD SELSYN CONTROL SYSTEM FOR ELEVATING JACKS IOHP 10 HP 10 HP RECEIVING -- JACK DRIVE SELSYNS POSITION 120 HP INDICATOR 120 HP 0-0 0------------J FIG. 14 s TD SQUIRREL CAGE MOTOR TRANSMITTING SELSYN - - - -. - - - - - .,.. -• --------• f POSITION SELECTOR r 00 J INSTRUMENT GEAR MOTOR J_ REFERENCE SYNCHRO • REMAINING 11 JACK DRIVES 3 PH 6Qrv 550VOLT EXCESS rWER LIMITS I a I • l 1 STARTER _1 DETECTOR STD SQ. CAGE DEMODULATOR GEAH HEAD MOTOR I I I LIFT I JACK "-----------t (') ,_ - r71 _J ::::./ lL...JTNSTRUMENT .L GEARING. SENSE SYNCHRO CLOSED LOOP CONTROL SYSTEM FOR ELEVATING JACKS FIG.15 1. I I I I I I I I. I I Ill • • • • PU PS P) SUPPORT COLUMNS (TYP) SECTION B-B (LOWER MAG ET CORE) L • If • • • .. • • .. '/AC H A K AG ET PATERN SHOWII IS LOCATED 0 LINES or EOUAL RADII. Al ER ATIV(l( USE PA ER fOR UPPER MAf1NET CORES IF rou CATION srnu(TURE IS SUITABLE TOTAL UM&R Of TIE RODS: 302 ( If ACUUM PUM 5) J.1/O 310 ( f 10 ,ACU PUHPS) ' A .. SECTIO A-A (UPPER MAG ET CORES) 1'1( ROD 5PACI JG • 3:i ~CM !> OR LE 5S • MAGNET SJPPORT POI S 0 Al UH BER Of n£ ROO!> • "06 r- -+---l_.l!:::=::::;:::;:_=_=_=~=r .... T-'-F_==r,====r~=====f ' 8 REVISION TRIU 11F LA (CUT C(CLOTRON OF TIE ROOS IIILI0ITII, SWIRD, ~ID MIii ASSOCIATES TOIIOIITO CA~A04 tt.-a '.a' • I' -o• .. 703 04 60000 I I I •• I I I I I I-I I I I I I I I LOtATIO or LHElllli6 JAClllt AMO HAGWCT llfTI 6 RODS SCA F NORHAl OP£RATI C. POS1TI01J A i< 1.1\.lOER VACUUH STOP ) DETAIL 'O' TIE-ROD ATTACHMENT g~LE"~l• FULL_S ze ILLUSTRATING Tll[ USE Of MECHANICAL STOPS TO P(RMIT INCAEA5(D Dlfl(CTION or ~UP1'0~T SfllUCTURE UNDER ELCVATIN6 OR OP£RATIMG COMDlflO><S WITllOllT OAl•V,GE TO COILS. EXTREME POSITION A K VEN ED OR ~TR Clll E ElEV.ATEO SfCTIOH 8-B r t1AC.lofT , ,_ --,_ o· l PR ARY STA urn S£'0 0.AR ,sr UC _, ,.. TA E -t· -I ' ! c 1~ SECTIO"-1 A-A SC•H 1~1• • l'•0' - -LEV ELLI NG ACKS $CME ,.. - r-0'--ERSIOH A 12 :AC...-s t5"0 s ,ERSIO B ll JA(KS 150 \IERSIOIII C 12 l;llOUPS 01 2NrO<!Au ( AHO OHf lfJC06• OD AS DRAWN - °::' REVISION (l(VA ., JA R005 I 'DIA I I -• B llVClllNG )AO<S ffRSION A J r: r, TRIUMF CYCLOTRON i~--1SLJPPORT STRUCTURES STRAIGHT BEA A CO CEPT Sllu\R U:NT~E Or I DETAIL E i;c•,r 5•-1•0• Oil WORTH. S£C1JIII, IIEAGID • ASSOCIATES COMSULTIMC { 1kttn TOROHTO YA"4COV'l'CR CA!IADA ...., AS SHOWN ., _,. 703 04 60001 I I . " " .. I .. I A ( I L A •;,. J.>1-4 I l -1-I I ,,. I ,.,. ,, I I l J ,0 .. ;I -I • 0 CD I • ' J 1-I TRIUMF CYCLOTRON DILIOITH, SECIIRD, MCASJD 11111 ASSOCIATES uwruo I SECTIO A-A CENTRE POST -TORONTO VAlolCOVvtR CANADA - .. _,, -- ·- 703 03 .
Report
eng
Unreviewed
Vancouver : University of British Columbia Library
Attribution-NonCommercial-NoDerivatives 4.0 International
Researcher
Conceptual Design Study for the support structure, elevating jacks and leveling jacks for a 500 MeV H‾ cyclotron
Text