P R A S I C ' 02. Simpozionul na:ional cu participare interna:ional; PRoiectarea ASIstat; de Calculator

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UNIVERITATEA TRANILVANIA DIN BRA.OV Catedra Design de Produs 0i Robotic2 impozionul na:ional cu participare interna:ional; PRoiectarea AIstat; de Calculator P R A I C ' 02 Vol. II Organe de ma0ini.

1 UNIVERITATEA TRANILVANIA DIN BRA.OV Catedra Design de Produs 0i Robotic2 impozionul na:ional cu participare interna:ional; PRoiectarea AIstat; de Calculator P R A I C ' 02 Vol. II Organe de ma0ini. Transmisii mecanice 78 Noiembrie Bra@ov, România IBN ON THE LITING JACK WITH CREW AND WORM GEAR Gheorghe MILOIU*, Radulorin MIRIC#**, George DOBRE** * C CONIND A Câmpina, ROMANIA **University POLITEHNICA of Bucharest, ROMANIA Abstract: The paper offers an introduction in the study of lifting jacks ith scre. In the first part, an analysis is made of the solutions offered by specialised companies, emphasising principal trends. The basis for the calculus of the loading limits is presented: mechanical, thermal and of buckling. The thermal calculus is presented for the transient operating mode (of short length). Useful data is extracted for the projection from the experience of some specialised companies. Keyords: lifting jacks, scre, orm gear, calculus, load acity. 1. Introduction Lifting mechanisms ith scre and orm gear are idely used in mechanical systems. The folloing main applications are notable: the stand positioning at the rolling mills building, the table positioning of the straightening machines for oodorking, the press; the lifting systems from the transport vehicles, variable geometry scenography in sho business, thermal and nuclear poer stations etc. Nevertheless, these mechanisms are rarely researched. The present paper approaches to main aspects: the construction optimisation and several basic calculation elements. 2. The construction of the mechanisms The lifting mechanisms are made of a orm gear, a scre transmission and a motor (or a geared motor) fixed ith an intermediate flange on the body of the orm gear. A coupling is placed beteen the driven shaft of the geared motor and the driving shaft of the orm gear. The construction is adapted for the assumption of some large axial forces in one direction or another. Principally, to types of lifting mechanisms are distinguished: mobile nut type or mobile scre type. The mechanisms of the ATLANTA Company are exemplified in the igure 1 [6]. or the mechanisms ith small axial clearances, a recent solution developed by TEMAG is presented in igure 2 [6]. The construction has rolling that assumes only axial forces or only radial ones, and is stiffer then the rolling for combining loads. A synthetic characterisation of the lifting mechanisms gears is given in Table 1, about: the housing construction, the of the orm, the of the orm heel and the assumption of the axial force of the driven scre, the systems for the adjusting of the clearance. or the driving shaft several existent solutions provide practically the same efficiency. But concerning the assumption of the axial force on the axis of orm heel, the system ith ball thrust is preferred, ensuring an efficient in the presence a certain prestress and having loer costs then other systems. Concerning the clearance

or the dimensioning of the motor (hich is fixed frontally on the housing of the orm gear) the speed of the driving shaft (motor) has to be knon.

2 278 adjusting in the system on the orm axis, the rigid ones are preferable to the others using adjustment by threaded parts. translation scre that assumes the axial load, the displacement of the scre/nut at a rotation of the driving shaft, the efficiency, the scre stroke. or the dimensioning of the motor (hich is fixed frontally on the housing of the orm gear) the speed of the driving shaft (motor) has to be knon. The scre lifting speed is v = s1 rot nm [ mm / min], (1) here: s 1 rot is the displacement of scre/nut (in in mm); n M the motor speed (in 1/min). Table 1 highlights the parameters of the series of mechanisms ALBERT [7]. The specialised companies offer a normal series and a slo one, hich differ in the speed ratio of the orm gear. pecial executions are encountered in applications of the rolling mills used in metallurgy. A A A A ig. 1. Usual solutions of lifting jacks ith scre and orm gear developed by ATLANTA [7] 3. The mechanism parameters Mechanisms for axial loads beteen 5 and 1000 kn ere developed. The basic parameters are: the maximal static load, the dimensions of the ig. 2. Recently solution of lifting jacks developed by TEMAG [7] 4. The calculus of the lifting jacks ith scre and orm gear The parameters of Table 2 are the result of a years of experience in manufacturing and service, and can be employed as point of reference in design.

3 Company Pfaff Albert Observations Duff Norton Radicon Costa Masnaga Atlanta temag 400 Benzlers Iuginas 150 cast cast cast cast cast cast Doubl ecast cast cast ro angular contact ball s s s s s Double ro angular contact ball s Deep groove ball + cylindrical ro angular contact ball s s 279 thrust s s Deep groove ball + cylindrical Needle + deep groove ball Cylindrical + journal By the collars of both s By the collars of both s By the rings upport disk at the of the one By adjustment disks By the collars of both s By the rings Table 1 A short characterisation of the orm gearbox solutions used in lifting jacks Clearance Clearance Bearing of Load acity [kn] ing driving shaft the driving the driven Hous Bearing of the adjustment of adjustment of the driven shaft shaft shaft troke max. 10 m peed max m/s Great axial stiffness The geometrical parameters of the orm gear. One of the folloing recommendations is used [3]: the numbers of teeth of the orm z 1 and heel z 2 : a z1 = (integer); (2) u z 2 = u z 1 (integer); (3) the exterior diameter of the orm according to the Table 3, here a is the centre distance of the gear. The driving torque is T2 T1 =, (4) u here the torque on the lifting scre axis is T 2 d2s = ax tg( s + s ) (5) 2 and: is the gear orm efficiency; u the gear ratio; s the angle of thread helices of the lifting scre; s the friction angle; d 2s medium diameter of the thread.

4 280 The sizes d 2 s, s and s are obtained using the thread parameters and a corresponded friction coefficient of the couple nutscre. The dependence of the driving torque on the scre axial force is given in igure 3. Axial force [kn] GT 5 GT 20 GT 30 here n1 is the rotational speed of the orm. A more accurate calculus of this angle is given by DIN [2]. The buckling of the lifting scre. The existence of the elastic or plastic buckling depends on mechanism construction, respectively the guide mode of the scre end. The calculation is made using the knon expression given by strength of materials [1]. The buckling load acity (for the size ith the nominal axial force in scre of 200 kn) in the scre length at the ALBERT mechanisms is given in igure 4. 2 GT Axial force [kn] GT 150 GT 200 a) GT 350 GT 300 GT 500 Torque [NMm] GT 1000 Maximal axial force 200 [kn] afety coefficient: Tetmajer 3 4 In Euler s domain 4 Type of mechanism: TG 200 The orm gear unit efficiency. Considering for discussion only the efficiency of the orm gear pair, this is given by Torque [NMm] b) ig. 3. The nut torque for obtaining the necessary axial force tg m =, (6) tg( + ) m here: m is the helical angle of the orm; the friction angle in the orm gear pair. That last angle is given approximately in the form [5], [3] tg0 tg =, (7) + 1+ n /(10 ) 1 1 The load acity of the orm gear pair. The calculation of load acity of the orm gear pair is normalised (as example by DIN 3996 [2]). There are four types of calculus by [2]: a) of the abrasive ear; b) of flank solicitation (to avoid the pitting); c) of orn tooth bending strength; d) of thermal safety. The abrasive ear calculus consists in the evaluation of the orn depth by: = J s, (8) cre length [mm] ig. 4. Buckling load acity as function of scre length (after [7]) m

5 281 here: J is the ear intensity; s m the friction road. This orn depth is compared ith the one limit ( lim ) given by typical expressions and based on experiment [2], resulting a safety factor lim = min. (9) The flank solicitation calculus (to avoid pitting) has the folloing steps: * effective flank pressure ( p m ) that depends of the geometry and materials of the orm and heel; medium flank pressure ( Hm ); limit value of the medium flank pressure ( HG ) ; safety factor in the form HG H = H min. (10) Hm The orn tooth shear strength calculus has the folloing steps: the effective shear stress ( ), determined by taking into account the eakening of the tooth by abrasive ear; limit shear stress ( G ); safety factor G = min. (11) The thermal calculus is given in [2] for to lubrication cases 1(splash and forced feed) and has the folloing steps: total poer loss composed by the friction poer in tooth gearing, in idle running, in s and in seals; sump temperature (for the case of splash lubrication; in the other case of the forced feed lubrication the oil temperature is controlled by cooling or flo); thermal safety factor, obtained by: temperature in the case of the splash lubrication T s lim = T min ; (12) poer in the case of the forced feed lubrication T PK = T min, (13) P V here: s lim and are the limit and, respectively, effective temperatures of the sump oil; PK and PV the cooling and, respectively, the total poer loss in the orm gearbox. It is orth mentioning that the transient operating conditions are not included in the DIN variant. The authors of the present papers approached the thermal calculus assuming a transient operating mode [4]. This calculus has the folloing steps valid for a small time increment t : calculation using the indications, typical values of constants and expressions given in DIN 3996 [2] of the total poer loss P V comprising friction poer in tooth gearing and s, neglecting the poer loss in idle running and in seals (very lo over a large number of calculations); this poer loss depends on the number of revolutions n hich varies in transient operating conditions; deriving the heat quantity given out in the little time interval t in the operating conditions Q = P, t ; 14) i V i calculation of the elements temperatures at the final time interval taking into account the specific heat and the mass of the components Qi1 + Qi i = + i1, mc ith m mass of the element, c specific heat of the material, i 1 initial temperature; determining the transferred heat quantity from an element to another at this final time increment ( t ) Q = ka t, (15) transf 1 2 i,1 i,2 here A is the heat transfer contact area; k the heat transfer factor; calculation of the ne temperatures of the elements after this heat transfer. The steps mentioned above are repeated (iterated) for the subsequent intervals t until the final orking cycle.

6 By applying this procedure, the variation in time of the instantaneous temperature of all components of the orm gear unit is compared ith the limit temperatures for each component. An example of this variation in time of the temperature of the housing is given in igure 5. One can see that housing temperature increases quickly in the beginning and after this it varies sloly. 76 [ o C] [min] t ig. 5. Evolution of the housing temperature during Conclusions The present paper discusses to main aspects regarding lifting jacks ith scre and orm gear: the construction optimisation and several basic calculation elements. 2. The construction is adapted to the assumption of some big axial forces in one direction or another. To types of lifting mechanisms are used for this: ith mobile nut or mobile scre. 3. The basic parameters are: the maximal static load, the dimensions of the translation scre that assumes the axial load, the displacement of the scre/nut at a rotation of the driving shaft, the efficiency, the scre stroke. 4. The basic calculation elements for this lifting mechanisms are: geometrical parameters of the orm gear; driving torque; orm gear unit efficiency; load acity of the orm gear pair: for abrasive ear; for flank solicitation (to avoid the pitting); for orn tooth bending strength; at thermal safety. 5. An on method for thermal safety in transient operating is succinctly discussed. References 1. Buzdugan, G. Proiectarea de rezisten./ în construc.ia de ma1ini. Editura Academiei Române, Bucure@ti, DIN Tragfähigkeitsberechnung von Zylinder chneckengetrieben mit Achseninkel O = 90 o, Miloiu, G, Constantin, T., Vintil;, H. Multicriterial selection of parameters for cylindrical orm gears. PRAIC 98, Vol. II, University Transilvania Bra@ov, 1998, p Miric;, R.., Dobre, G., Miloiu, G. Thermal calculus of the orm gearboxes in the transient operating conditions. cientific Bulletin of North University of Baia Mare, erie C, Volume XVI, ascicle: Tribology, Machine Manufacturing Technology (Papers of the International meeting of the Carpathian Region pecialists in the ield of Gears, 4 th Edition), Baia Mare, 2002, p Moisan, J. M. Considérations intéressantes sur le rendement des engrenages à vis tangentes bien réalisés. Congres Mondial des Engrenages, Paris, 1977, vol. 2, p Niemann, G., Winter, H. Maschinenelemente, Vol. III, pringer Verlag, Technical documentation from the specialised companies: ALBERT Maschinenfabrik Vöcklabruck, ZIMM Maschinenelemente Bregenz, ENZELDER Mech. Werkstäte, TEMAG teirische Maschinen und Geräteban Austria. GC Maschinenfabrik Cosig, PA silberblan Hebezeugfabrik Augsburg, ATLANTA Zahnradfabrik Bietigheim Bissingen Germany. DU NORTON London, DAVID BROWN RADICON Great Britain. M MOLDAVKE TROJARNE Moldava nad Bodvon lovakia. MARTINETTI MECCANICI Castel Maggiore, COTAMANAGA Como, ERVOMECH Anzola Emilia Italy. BENZLER Helsingborg eden. IUG Craiova, NEPTUN Câmpina Romania.

THERMAL CALCULUS OF THE WORM GEARBOXES IN THE TRANSIENT OPERATING CONDITIONS

THERMAL CALCULUS OF THE WORM GEARBOXES IN THE TRANSIENT OPERATING CONDITIONS Radu Florin MIRICĂ*, George DOBRE*, Gheorghe MILOIU** * University POLITEHNICA of Bucharest, Department of Machine Elements