El I II I )111 II El

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1 THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS 345 E. 47th St, New York, N.Y GT-104 The Society shall not be responsible for statements or opinions advanced impapers or thicussion al meetings of the Society or of its Divisions or Sections, or printed in its publications. Discussion is printed only lithe paper is published in an ASME Journal. Authorization to photocopy material for Internal or personal use under circumstance not falling within the fair use -provisions at the Copyright Act is granted by ASME to libraries and other users registered with the Copyright Clearance Center (CCC) Transactional Reporting Service provided that the base tee of $0.30 per page is paid directly to the CCC, 27 Congress Street Salem MA Requests for special pennission or bulk reproduction should be addressed to the ASME Technical Punishing Department Copyright by ASME All Rights Resented Printed in U.S.A TOUCHDOWN BEARING DEVELOPMENT FOR A MAGNETIC BEARING SYSTEM USED IN A HIGH TEMPERATURE GAS TURBINE El I II I )111 II El Francis P. Marchand, Jr. The Charles Stark Draper Laboratory, Inc. 555 Technology Square Cambridge, Massachusetts Telephone: (617) FAX: (617) fmarchand@draper.com ABSTRACT This papers describes the design and development of a touchdown bearing system. The touchdown bearing is one of the important elements in any high speed magnetic bearing system. The touchdown bearing functions to fully off-load or load share with the magnetic bearing and prevent catastrophic system failures. The proper design of this component is a critical step in the design of the total system and was not treated as a stand-alone sub-component. The impact loads which these bearings have to endure is a challenge to the designer. The high temperature environment on this bearing adds additional constraints to the design options which the designer has to work with. A design was produced, which will function under these conditions. The use of silicon nitride balls with sliver coated CPM Rex 20 races will be described. The bearing is mounted on a replaceable journal, which in turn is mounted on the magnetic bearing. Design and test data for use of these bearing materials in a high temperature gas turbine application is included in this paper. NOMENCLATURE DB back-to-back preload Diam diameter ON speed factor (bearing pitch diameter (D) in millimeters times rotational speed (N) in RPM) GPa gigapascal MPa megapascal OD outside diameter Rc Rockwell C RPM revolutions per minute RT room temperature Si 3N4 silicon nitride g/cc grams (mass)/cubic centimeter lcn kilonewton mm millimeter nun2 square milimeter ppm parts per million C degrees Celsius OVERVIEW A touchdown bearing system has been designed to meet the force, environment, and volume requirements for the turbine engine system. The application's environment imposed severe design constraints on the overall selection of materials and bearing size. The touchdown bearing is required to operate at up to 482 C and rotate at up to 22,000 RPM or 4.6 million DN. ON values higher than 2 million are considered to exceed the state-of-the-art in bearing technology. Liquid bearing lubricants will not be available in the intended application. Alternate methods of lubrication must be developed and utilized to meet the life requirements of this application. This paper will describe one such method, currently in use in very high temperature applications. The bearing's design arose from repetitive iterations of computer-aided analysis for thermal-induced stresses, Von Mises stresses, and bearing raceway stresses and deflections. A common touchdown ball-bearing design is used in both the front and aft turbine engine locations (Reference I). The front touchdown bearing handles both radial and axial loads while the aft bearing handles only radial loads. The bearings' inner rings will be mounted on a replaceable journal, which is designed to clamp the inner rings together and act as a wear surface. DESIGN GOALS AND REQUIREMENTS A design goal was established for the touchdown bearing to be capable of assembly integration with the magnetic actuator bearing. The of the touchdown bearing's outer ring will be bolted to the engine frame or the magnetic actuator bearing. The bearing will be stationary (i.e. non-rotating) until a touchdown event occurs, at which point the inner rings will rotate. The bore of the inner journal will have sufficient radial clearance, such that the rotor's shaft will be allowed to freely move axially through the bore at all times. The touchdown bearings shall be capable of surviving multiple touchdown events as outlined in Table I, below. The G loading is based on the weight of the turbine engine rotor and all of ancillary components attached to the rotor. The touchdown events in Table I will affect the operational life Downloaded From: Presented at the International 11/03/2018 Gas Terms Turbine of Use: & Aeroettgine Congress & Exhibition Orlando, Florida June 2 June 5,1997

2 of the touchdown bearing, where a minimum time limit has been established. Table I Touchdown Events Condition G RPM Engine Time Limit Thrust Transportation & none ground handling Ground windmilling none In-flight windmilling 1.5 5, minutes In-flight shutdown 2 ma max. 30 seconds The published data for the selected CPM Rex 20 race material did not include information on tensile or yield strength. Telephone conversations with a bearing manufacturer, who uses.cpm Rex 20 material for bearings (Reference 4), did not result in strength values for the material. A telephone communication was held with the steel manufacturer, who stated "the CPM Rex 20 has similar tensile strength properties as M50 at the same relative hardness" (Reference 5). Since it was previously established that the inner races will rotate at very high speeds and result in large hoop stresses, tensile strength becomes a critical design feature. Strength tests were performed on actual bearing inner races, which replicate the shape of the proposed design. Figure 1 shows the shape and half cross-sectional area of the race tensile test piece, used during the strength evaluations. Additional forces acting on the bearings are due to rotor dynamics from high-g maneuvering, rotor imbalances and control simulations. The touchdown bearing's final operational requirements were derived from rotor dynamic evaluations and controls simulations by others (References 2 and 3) and are summarized below: Operating temperature range: -40 C to 482 C Speed range: Touchdown Gap: 13,400 RPM to 22,000 RPM mm radial and/or axial. Radial bearing peak force: kn front bearing, IcN aft bearing Axial bearing peak force: Lubrication: lcn forward or kn rearward on front bearing Self contained within the bearing MATERIALS SELECTION Two essential design requirements limited the choice of materials, which could be used for the touchdown bearing. These are the very high operating temperature (up to 482 C) and fast rotational speed (up to 22,000 RPM) for the bearing's size. Any material selected must have a high hot hardness, wear resistance, and relatively high tensile strength. The tensile strength is the speed limiting factor in bearing material selection. The following bearing components were interactively studied for material selection: I) inner and outer races, 2) rolling elements (balls), 3) mounting journal, 4) lubrication. Inner And Outer Races Conventional bearing steels were evaluated for the races, but abandoned because of their reduction in hardness at operating temperature and low tensile strength. These materials included 440C stainless steel, chrome steel and M50 tool steel. The selected race material for evaluation has seen extensive use in bearings for medical X-ray tube applications. An X-ray tube is run in a hard vacuum, without liquid lubrication, at temperatures of 538 C and at a high, instantaneous rotational speeds. This material is known as CPM Rex 20. The high rotational bearing speeds will generate large centrifugal ball loads, which can be minimized by selecting a lower density (than steel) ball material. Silicon nitride ceramic material has been utilized in rolling element bearings for this feature and was evaluated for this application. Figure 1 CPM REX 20 Tensile Test Specimen Four different sample preparation conditions were performed on the specimens prior to the tensile strength tests. All tests were performed at room temperature. The thermal conditions were: I. As received. 2. Heat cycled to 593 C and back to room temperature three times. 3. Heat cycled to 593 C and back to room temperature three times, then heat cycled to 760 C and back to room temperature three times. 4. Heat cycled to 760 C and back to room temperature three times. The results of the tests are tabulated in Table IL The results show the "As Received" races and those heated to 593 C have the same value for 0.2% yield strength and ultimate tensile strength, indicating parts are brittle, as would be expected for these types of materials. Overall, the strength of the parts subjected to the 593 C thermal cycle decreased compared to the "As Received" races. The hardness also decreased slightly. This is an indication the 593 C preparation temperature approached or exceeded the tempering temperature, which is typically performed at 551 C, but unknown for these test samples. There was a marked increase in material strength when the races were.subjected to the 760 C thermal cycle. There was also a separation between the yield strength and ultimate tensile strength, which is an indication of being more ductile. The hardness also had an appreciable decrease, indicating the parts were fully annealed. It can be concluded if the component temperature can be kept below approximately 593 C, then the yield strength will be approximately 731 MPa. This value will be used as a figure of merit for overall bearing design. 1 2

3 Table II CPM REX 20 Material Properties Thermal Condition (See text) Average No. of Samples Time at Temp (hours) X-Sectional area (mm 2) Fracture (IcN) Yield (kn) Ultimate Tensile Strength (MPa) % Yield Strength (Mpa) Rockwell C Hardness Diameter Before Cycling (mm) nm Diameter After Cycling (mm) nm Width Before Cycling (mm) nm Width After Cycling (mm) nm nm= not measured. Rolling Elements The high rotational bearing speeds will generate large centrifugal ball loads, which can be minimized by selecting a lower density (than steel) ball material. Silicon nitride ceramic material has been utilized in rolling element bearings for this feature in turbine engines and was evaluated for this application. Silicon nitride also has a maximum use temperature of 1,000 C and Hertz compressive strength of 28 Gpa (Reference 6). The main disadvantage of silicon nitride is its relatively low thermal expansion coefficient compared to bearing steels. Silicon nitride material properties are shown in Table III. Mounting Journal The touchdown bearing design was selected to be a preloaded pair, angular contact ball bearing for rotational stability. This design necessitates the need for a mounting housing for either the outer rings, inner rings or both. The material for the mounting housing requires a careful match-up for the coefficient of thermal expansion to the bearing rings. In the ease of an inner ring shaft or journal housing, the modulus of elasticity, yield strength, hardness and wear resistance are also important design features. Draper selected an ultra-high-strength maraging steel, identified as Vasco Max 350 CVM and manufactured by Teledyne VASCO for evaluation in this application. Lubrication The sponsor of this design investigation pre-determined conventional liquid lubrication would not be available. Silver plating was selected for the inner and out raceways, the bore and external faces of the mounting journal and for the bearing's balls. The silver plating will be the only means of reducing the coefficient of friction between the rolling or sliding components. As a comparison, the X-ray tube bearing design uses a thin film silver plating on the raceway surfaces but operates in a vacuum environment. As noted earlier, the X-ray tube bearing must operate at temperatures of 538 C. A turbine engine application operates in air so high temperature oxidation is a concern for any materials selection. Many turbine engine bearings use a silver plated ball separator for emergency back-up lubrication. The operating temperature of the silver plated ball separator is similar to the proposed touchdown bearing. Reference 7 indicates silver in the presence of air at temperatures of 600 to 800 C did not show any appreciable attack or tarnish. Lubrication will be the life limiting factor. - Additional testing will be required to substantiate any analytical life calculations. MATERIAL PROPERTIES The published properties of the three materials identified (CPM Rex 20, silicon nitride and Vasco Max 350 CVM) are presented in Table III. The table compares the properties against M50 VIM-VAR bearing steel. It should be noted, Draper's hardness measurement on CPM Rex 20 was very similar to the published data (Rc 67 vs. 67.5). Table III Touchdown Material Properties Property M50 1 CPM Rex 202 VascoMax Silicon 350 Nitride CVM 3 Density (g/cc) RT Young's Modulus (GPa) Young's Modulus at 315 C (GPa) Young's Modulus at 538 C (GPa) RT Thermal Expansion (ppinrc) C Thermal Expansion (ppm/ C) RT Yield Strength (MPa) C Yield Strength (MPa) 1379 RT Hardness (Rc) > C Hardness (Rc) Notes: I. Reference 8., 2. Re erence 9., 3. Reference 10., 4. Reference 6., 5. Estimated from M50 data. 6. Refe ence 11. STRESS AND LIFE ANALYSIS Draper performed a stress and life analysis on the touchdown bearing design configurations, based on the materials selected and evaluated in the previous section. The personal computer based software, A. B. Jones High Speed Ball and Roller Analysis Program, Version No , was used for the analysis. The following assumptions were used in performing the analysis: I mm diameter Si 3N4 balls per race. 2. pitch diameter of mm initial contact angle kn DB (back-to-back) preload. 5. Rotational speed of 22,000 RPM. 6. Young's Modulus for Rex 20 is 234 GPa at room temperature. 7. Young's Modulus for Rex 20 is estimated (from Table III) as 171 GPa at 315 C and 123 GPa at 538 C. 8. Maximum Hertz compressive strength of S1 3 N4 material is 28 Gpa (Reference 6). 9. Maximum Hertz compressive strength of bearing steel is 4.0 Gpa at the on-start of raceway damage (brinelling). 3

4 The results of the analysis are found in Table IV, below. The maximum Hertz stress was calculated as 2.10 GPa on the front bearing, when the combined maximum radial and axial loads were used at operating temperature. The room temperature stresses increased in all cases when the temperature related material properties were substituted. The L-I0 life was calculated for both the front and rear bearings under maximum operational loads, temperature and rotational speed. Harris (1984) indicates L-10 is the fatigue life that 90% of the bearing population will endure. The fatigue life is a time dependent function (rotational speed) estimated by the ratio of the basic dynamic capacity of the bearing and the normal ball load to the cubed power (Reference 12). L-10 life may include modification factors for material cleanliness and hardness, lubrication and others. The results indicate the front bearing will have an estimated L-10 life of 9.9 hours and the rear bearing 24.8 hours. These values are well within the design requirements for in-flight windrnilling of 30 minutes and in-flight shutdown of 30 seconds. As a design exercise, the loads used on the front bearing were increased by a factor of five, to determine the upper design limitations. This resulted in a maximum Hertz stress of 3.78 GPa with room temperature conditions. This stress value is still within acceptable limits for bearing rated steels, although the theoretical L-10 life is only 0.14 hours. s of this magnitude may occur briefly in the circumstance of abrupt magnetic bearing shutdown. Table W Touchdown s vs. Stresses Thermal Condition Radial (in) Axial (kn) Maximum Hertz Stress (GPa) Max. Contact (IN) L-10 Life (Hours) RI Front 315 C RT Aft RI C THERMAL AND INERTIAL ANALYSIS The effects of high operating temperature and high rotational speeds were evaluated in order to complete the bearing design. Because the front and rear bearings operate at two different temperatures, it is anticipated there will subtle size adjustments made to each bearing, The critical dimensions, evaluated for thermal effects, are joumal OD to bearing bore fit, journal bore diameter, ball diameter size change and bearing pitch diameter change. These are depicted in Fig. 2. The journal bore diameter and bearing pitch diameter are given additional considerations due to inertial effects. The thermal growths were calculated using the material properties found in Table III and bearing dimensions noted in Fig. 3. These values are tabulated in Table V, below. The diametral growth due to inertial effects are also found in Table V. The journal bore diameter increase was derived from ANSYS Ver. 5.2 and the pitch diameter increase was derived using the Jones Analysis Program. Journal Outer Race Journal to inner race interference fit Journal to rotor shaft clearance (.203 tom) Rotor Shalt Figure 2 Touchdown Fits And Clearances Inner Race Bore The race bore will increase in size faster than the journal OD due to the slight mismatch of the coefficient of thermal expansion between the journal material and the race material (6.3 ppm vs ppm. respectively). Table V shows this effect as the delta inner race bore to journal OD data. If both bearings are initially mounted onto the journal "line-to-line" then the front bearing will result in a mm clearance and the rear bearing a mm clearance. To overcome this effect, both bearings will be mounted with a room temperature interference fit equal to the operating temperature clearance value. The effects of an interference fit will result in a tensile stress on the bore of the inner race. The stress was calculated using ANSYS and found to be approximately MPa. Mounting Journal Bore The journal bore diameter will increase in size due to both thermal and inertial effects. The operating clearance between the journal bore and rotor shaft OD must be approximately mm to mm radially during operating conditions for touchdown purposes. The manufactured size of the journal bore will be compensated for, based on rotor material properties to allow for this operating clearance. Table V gives the thermal increase of mm and mm on diameter at 315 C and 482 C, respectively. The ANSYS calculations indicate the bore will have an additional increase in size by approximately mm at 22,000 RPM. The effects of the high rotational speed will result in a tensile stress on the bore of the inner race. The amount of the stress was calculated using ANSYS and found to be approximately 445 MPa with the bearing having an initial mm interference fit. This stress level is well within the measured yield strength of the material (731 MPa) (Reference 11). Rolling Elements The design of the bearing relies on the use of a full complement of balls. Typically, the bearing is designed, such that there is a total ball gap between the first and last ball of approximately one-half ball diameter at room temperature and nominal design conditions. The ball gap will vary considerably due to ball and raceway manufacturing tolerances and ultimately due to thermal and inertial effects. The goal is a full ball complement design so that there is always a ball gap during all operating conditions, which will minimize ball-to-ball loading. Typical manufacturing tolerances were used to determine the change in ball gap. For the ball diameter it was assumed to be ±

5 ' mm and for the pitch diameter it was assumed to be ± mm. The full range of ball gap is presented in Table VI. The ball diameter and pitch diameter were varied by the above tolerances and also due to the thermal and inertial effects presented in Table V. The gap was calculated using equations found in Reference 13. The nominal conditions are 21 C, 12.7 mm ball diameter and mm pitch diameter. This resulted in a ball gap of 6.33 nun (approximately onehalf ball diameter). The minimum ball gap (4.22 mm) occurs at -40 C and the largest ball diameter and the largest ball gap (15.04 mm at 482 C with the smallest ball diameter. This last value indicates the ball gap has increased to slightly more than one full ball. From the data, it can be seen the ball gap will inherently become larger because of the increasing pitch diameter and decreasing ball diameter of silicon nitride compared with CPM Rex 20 race material.. The relative size change in ball diameter in relation to the race diameters indicates the balls may become loose in the bearing at operating temperature. The effect of this will eliminate or decrease the bearing's preload. This will result in excessive ball skidding and premature wear. Table V shows the front bearing has balls mm smaller than its race and the aft bearing are smaller. In order to compensate for this effect, the bearings will be assembled at room temperature with oversized balls, which are proportional to their expected size change at operating temperatures. DESIGN CONFIGURATION The touchdown bearing design configuration is shown in Fig. 3. The bearing uses a back-to-back duplex configuration with a flanged, integral outer race (two ball rows in one race), two separate inner races and mm diameter balls per race (full ball complement). After the preload is established by grinding or size matching, the inner races are clamped together by an interference fitted journal and lock nut. Nominally, the bearing OD (excluding the flange) is 240 mm and the journal bore is 173 mm. The width of the outer race is approximately 49 mm and the width across the journal and nut is approximately 65 mm. The outer race is flanged with 12 scallops and 12 mounting holes. The scallops are used to allow the magnetic bearing actuator connectors to pass through. The bearing design will have 52% inner and outer race curvatures, a 25 initial contact angle, a 1.11 kn preload (DB), and an mm pitch diameter. Outer Race Mounting Flange Inner Race Preload Nut Table V Touchdown Thermal & Inertial Properties Property Either (at 21 C) Front (at 315 C) Aft (at 482 C) Thermal Related Changes Journal Bore Increase N/A Inner Race Bore Increase N/A Journal OD Increase N/A Delta Inner Race Bore to N/A Journal OD Pitch Diameter Increase N/A Ball Diameter Change (Ref to CPM Rex 20) N/A Inertial Related Chan es Journal Bore Increase Pitch Diameter Increase ' Notes: Dimensions are in mm. N/A= not applicable. Temp ( C) Table VI Touchdown Ball Gap Data Ball Diam (mm) Pitch Din (mm) Number of Balls Ball Gap w/52 Balls (mm) Figure 3 Touchdown Design The materials selected for the bearing are CPM Rex 20 race material, which will be heat treated according to the manufacturer's recommendations. The balls will be silicon nitride, manufactured to CERBEC's NBD-200 material specification and finished to ABMA Grade 10 requirements or better. The journal and nut material will be Vasco Max 350 CVM and heat treated, according to the manufacturer's recommendations. All contacting, rubbing, or wear surfaces will be silver coated to aid in the lubricity of the bearing. SUMMARY Based on the design requirements of loads, rotational speeds, operating temperature and space considerations, a touchdown bearing has been designed to meet those requirements. A single bearing design will be used for both the front and rear locations. Each bearing will be adjusted for thermal effects on the race to journal interference fit and ball size. The front bearing is capable of handling both radial and axial loads and the rear bearing the specified axial loads. Stress analysis of the bearing design indicates the bearing is capable of handling loads of up to five times the specified design requirements. 5

6 REFERENCES I. Kelleher, W. P. & Kondoleon, A. S., 1997, "A Magnetic Suspension System for High Temperature Gas Turbine Applications, Part HI; Magnetic Actuator Development", Submitted concurrently with this paper. 2. Antikowiak, B. M. & Nelson, F. C., 1997, "Rotordynamic Modeling for an Actively Controlled Magnetic Gas Turbine Engine, Pan I", Submitted concurrently with this paper. 3. Scholten James, 1997, "A Magnetic Suspension System for High Temperature Gas Turbine Applications, Pan II; Control Systems Design", Submitted concurrently with this paper. 4. Hanson, Robert, January 17, 1996, Personal Communication, Miniature Precision. 5. Taney, Edward, January 23, 1996, Personal Communication, Crucible Specialty Metals. 6. CERBEC, TECHNICAL DATA SHEET #1A, CERBEC, Ceramic Company 7. Uhlig, H. H., 1948, The Corrosion Handbook, John Wiley & Sons, New York pp Alloy Digest, Filing Code: TS-415, Alloy Digest, Inc., July Alloy Digest, Filing Code: TS-398, Alloy Digest, Inc., May Alloy Digest. Filing Code: SA-251, Alloy Digest, Inc., May Standley, P., April 23, 1996, CPM Rex 20 Tensile Testing and Hardness, Draper Laboratory Internal Memorandum (E ). 12. Harris, T. A., 1984, Rolling Analysis, John Wiley & Sons, New York pp New Departure, 1946, "Analysis of Stresses and Deflections", Vol. 1, pp 8-9,. 6