行星架的數(shù)控加工與選用設計【NGW72-16二級行星圓柱齒輪減速器】

行星架的數(shù)控加工與選用設計【NGW72-16二級行星圓柱齒輪減速器】,NGW72-16二級行星圓柱齒輪減速器,行星架的數(shù)控加工與選用設計【NGW72-16二級行星圓柱齒輪減速器】,行星,數(shù)控,加工,選用,設計,ngw72,16,二級,圓柱齒輪,減速器 MMK TRITA-MMK 2005:01 ISSN 1400-1179 ISRN/KTH/MMK/R-05/01-SE Relations between size and gear ratio in spur and planetary gear trains by Fredrik Roos planetary gears are commonly known to be compact and to have low inertia. Keywords Spur Gears, Planetary Gears, Gearhead, Servo Drive, Optimization Contents 1 INTRODUCTION / BACKGROUND. 5 2 EQUIVALENT LOAD. 6 3 SPUR GEAR ANALYSIS. 8 3.1 GEOMETRY, MASS AND INERTIA OF SPUR GEARS. 8 3.1.1 Geometrical relationships . 8 3.1.2 Gear pair mass . 9 3.1.3 Inertia . 9 3.2 NECESSARY GEAR SIZE . 10 3.2.1 Hertzian pressure on the teeth flanks . 10 3.2.2 Bending stress in the teeth roots. 12 3.2.3 Maximum allowed stress and pressure. 13 3.3 RESULTS AND SIZING EXAMPLES. 14 3.3.1 Necessary Size/Volume. 14 3.3.2 Gear pair mass, geometry and inertia. 15 4 ANALYSIS OF THREE-WHEEL PLANETARY GEAR TRAINS . 19 4.1 GEAR RATIO, RING RADIUS, MASS, INERTIA AND PERIPHERAL FORCE. 19 4.1.1 Gear ratio and geometry . 19 4.1.2 Weight/Mass . 20 4.1.3 Inertia . 20 4.1.4 Forces and torques . 22 4.2 PLANETARY GEAR SIZING MODELS BASED ON SS1863 AND SS1871. 23 4.2.1 Sun planet gear pair . 23 4.2.2 Planet and ring gear pair . 24 4.2.3 Maximum allowed stress and pressure. 26 4.3 RESULTS AND SIZING EXAMPLES. 27 4.3.1 Necessary size/volumes . 27 4.3.2 Weight, radius and inertia. 28 5 COMPARISON BETWEEN PLANETARY AND SPUR GEAR TRAINS . 32 6 CONCLUSIONS. 34 7 REFERENCES . 35 Relations between size and gear ratio in spur and planetary gear trains 4(35) Nomenclature Pressure angle rad Helix angel ( = 0 for spur gears) Transverse contact ratio Load angle rad Poissons number Mass density kg/m 3 F Root bending stress Pa Fmax Maximum allowed bending stress Pa H Hertzian flank pressure Pa Hmax Maximum allowed hertzian flank pressure Pa Angular velocity rad/s b Gear width m b c Carrier width m d Gear reference diameter m d a Gear tip diameter m d b Gear base diameter m E Module of elasticity Pa F Force N J Mass moment of inertia kgm 2 K F Factor describing the division of load between teeth K F Load distribution factor for bending K H Factor describing the division of load between teeth K H Load distribution factor for Hertzian pressure k ro Relation between outer and reference diameter of ring gear m Module m M Mass kg n Gear ratio of a complete gear train ( in / out ) p b Base pitch r Gear reference radius m S Safety factor T Torque Nm u Gear ratio of a single gear pair v Peripheral velocity m/s Y F Form factor for bending Y Helix angle factor for bending Y Contact ratio factor for bending z Number of teeth Z H Form factor for Hertzian pressure Z M Material factor for Hertzian pressure Z Contact ratio factor for Hertzian pressure Index 1 The small wheel (pinion) of a gear pair Index 2 The large wheel (gear wheel) of a gear pair Index r The ring gear in a planetary gear train Index s The sun gear in a planetary gear train Index p The planet gears in a planetary gear train Index c The planet carrier in a planetary gear train Relations between size and gear ratio in spur and planetary gear trains 5(35) 1 Introduction / Background This work was initiated within a research project about design and optimization methods for mechatronic systems. The goal with that research project is to derive methods for optimization of mechatronic actuation modules, with respect to weight, size and/or efficiency (Roos the heat generated in the motors winding is given by the RMS value of the motor current. Since the current is proportional to the motor torque, the RMS torque may be used for motor dimensioning. 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 1000 500 0 500 1000 time s Load profile, required torque and position Ang. Velocity rad/s Angle rad 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 2 40 20 0 20 40 60 RMS RMC Max time s torque Nm Figure 2. Example of an inertial load cycle, with RMS, RMC and max norms shown. Gear design is traditionally focused on strength of the gears. The load on the gear teeth is cyclic and therefore gear failure is most often a result of mechanical fatigue. The two classical limiting factors in gear design are surface fatigue and tooth root bending fatigue. The combination of cyclic loading of the gear teeth when in mesh and an applied load that varies with time makes it more difficult to find an expression for the equivalent load, than in the motor case. The exponent used in the torque norms used for gear sizing is not 2 as in the RMS norm, but ranges from 3 to 50 (Anthony 2003). These expressions for equivalent load are based on the so-called linear cumulative damage rule (the Palmgren-Minor rule). It assumes that the total life of a mechanical product can be estimated by adding up the percentage of life consumed by each stress cycle. The number of stress cycles on each tooth in a gear train can be huge during a lifetime. Anthony (2003) exemplifies this with a three-wheel planetary gear in which one sun gear tooth will be exposed to almost 3 million load cycles during an 8-hour period at 2000 rpm. Relations between size and gear ratio in spur and planetary gear trains 7(35) Area of unlimited load cycles Stres s l oad (l og S) Load Cycles (log N) E = 3 for bearings E = 8.7 for case hardening steel E = 17 for nitrided steel 10 6 10 9 10 3 10 12 E = 84 for carbo nitrided steel Figure 3. Whler curves for different steels. The exponent to use in the calculation of equivalent load depends on material type, heat treatments, and loading type (Antony 2003). It is however not obvious that the Palmgren-Minor rule can be used for infinite life design ( 10 6 load cycles), especially not in an application where the teeth will be subjected to the peak load more than 10 6 times. In fact only in applications where the total number of load cycles is below 210 6 is a higher load than the endurance limit load permissible (Antony 2003). This means that for infinite life dimensioning, the gears should be dimensioned with respect to the peak torque in the load cycle. Of course there are exceptions to this, for example load cycles where the peak load occurs while the gears are standing still. The calculated equivalent continuous torque, T cal is hence, for unlimited life design given by: max )(tTT cal = (1) This is the approach taken in this report; it is assumed that the teeth are subjected to the peak load more than 10 6 times, and therefore is the peak torque used for dimensioning. This research area is however very complex and are not investigated further in this report. By this approach, equation (1), is at least not a too low equivalent torque used. The sizing procedure gets even more complicated when the bearings are considered. For bearings, the Root Mean Cube (RMC) value of the load is often used as the equivalent continuous load (comp. Figure 3). This report will however only treat the actual dimensioning of the gears, not the bearings. But it should be noted that it may be the bearings that limit the maximum gear load. Relations between size and gear ratio in spur and planetary gear trains 8(35) 3 Spur gear analysis The analysis made here is mainly based on the formulas presented in the Swedish standard for calculation of load capacity of spur and helical gears, SS1871 and standard SS1863 for the spur gear geometry. Figure 4 shows a spur gear, to simplify the analysis only spur gears with no addendum modification are treated. Figure 4. Spur gear. 3.1 Geometry, mass and inertia of spur gears. 3.1.1 Geometrical relationships In order to simplify the rest of the analysis, it is useful to derive some simple geometrical relations. The gear ratio u is defined as: 2 1 1 2 1 2 = z z r r r r u in out (2) The center distance, a between the wheels is given by: 1221 rarrra =+= (3) Combining equations (2) and (3) gives: 1 1 r ra u = (4) Pinion Gear Wheel T in T out a r 2 r 1 Relations between size and gear ratio in spur and planetary gear trains 9(35) Finally, combining equations (3) and (4) gives the expressions for r 1 and r 2 1 1 + = u a r (5) 11 22 + = + = u au r u a ra (6) 3.1.2 Gear pair mass A gear wheel is here modeled as a cylinder, an approximation that is quite accurate. The mass, M of one gear is hence given by: 2 brM = (7) Where b is the face width, r is the reference radius and is the mass density of the wheel. The total mass of a gear pair can be expressed as: )( 2 2 2 121 rrbMMM tot +=+= (8) Finally by combining equation (2), (5) and (8) the following expression for the gear pair mass is obtained: 2 2 222 1 )1( 1 )1( u u baubrM tot + + =+= (9) 3.1.3 Inertia The inertia, J of a rotating cylinder is given by: 2 2 cyl cylcyl r MJ = (10) The inertia reflected on the pinion shaft (axis 1) of a gear pair is hence given by: 2 2 22 2 2 12 1 2 2 2 2 2 1 1 2 2 1 2 2 2 2 u r br r br u r M r M u J JJ tot +=+=+= (11) Figure 5. Gear mesh p F p F Relations between size and gear ratio in spur and planetary gear trains 10(35) Which, if combined with equations (5) and (6) result in the following expression of the gear pair inertia: 4 2 4 4 24 4 4 )1( 1 2 )1()1( 2 u u ba u ua u ab J tot + + = + + + = (12) 3.2 Necessary gear size According to SS 1871, the necessary gear size is determined from the teeth flank Hertzian pressure and the teeth root stress. Neglecting losses such as friction the peripheral force, see figure 5, acting on a gear tooth is given by: au uT r T F out out out p )1( + = (13) For a given load the necessary gear size is determined as function of gear ratio and number of pinion teeth. Depending on material properties, gear ratio, number of teeth, etc., either the flank pressure or the root stress sets the limit on the gear size, in general both stress levels must be checked. 3.2.1 Hertzian pressure on the teeth flanks The Hertzian pressure on a teeth flank is given by (SS1871): ubd uKKF ZZZ HHcal MHH 1 )1( + = (14) For gears with no addendum modification the form factor Z H is given by: t b H Z 2sin cos2 = (15) As shown below, the transverse section pressure angle t is, for spur gears, the same as the normal section pressure angle, n . The pressure angle will therefore only be noted as from now on. = nt n t 0 cos tan tan (16) Since the helix angle is zero on the pitch cylinder () it will be zero on the base cylinder too ( b ) : 1cos0 cos coscos cos = bb (17) This leads to the following expression for Z H : 2sin 2 = H Z (18) The material factor Z M is given by (SS1857): + = 2 2 2 1 2 1 11 2 EE Z M (19) Relations between size and gear ratio in spur and planetary gear trains 11(35) Where E is the module of elasticity of the respective gear and is poissons number. For spur gears the contact factor, Z is according to SS 1871 given by: 3 4 =Z (20) Where is the contact ratio. For an external spur gear pair it is, according to SS1863, given by: + = ww baba b a dddd p sin 22 1 2 2 2 2 2 1 2 1 (21) Where a w is the center distance between the wheels, for gears with no addendum modification, aa w = . The base pitch, p b is given by: cosmp b = (22) Where m is the module, which is defined as: 12 2 1 1 )1( 2 zu a z d z d m + = (23) The tip and base diameters, d a and d b, are, for external spur gears, given by (SS1863): mdd a 2+= cosdd b = ) 44 (sin) 44 cos1( cos44cos)2( 2 22 2 2222 222222222 z z d z z ddd z d mddmmddmddd ba ba +=+= =+=+= (24) From equation (5) and (6) the following expressions for the gears diameters can be derived: 1 2 , 1 2 21 + = + = u au d u a d (25) By combining equations (24) and (25), and inserting them into equation (21), the following expression for the contact ratio is obtained: + + + + + = sin2) 44 (sin )1( 4 ) 44 (sin )1( 4 2 1 cos2 )1( 22 1 1 2 2 22 2 1 1 2 2 2 1 a uz uz u ua z z u a a zu += sin)1( 44 sin 44 sin cos2 22 1 1 2 2 1 1 21 u uz uz u z z z (26) Inserting equation (13) and (25) into equation (14) results in: 22 3 2222 2 )1( uba uT KKZZZ cal HHMHH + = This equation can be rewritten as: 2 max 2 3 2222 2 )1( H cal HHMH u uT KKZZZba + = (27) Relations between size and gear ratio in spur and planetary gear trains 12(35) Where Z H , Z M , Z , are given by equation (18), (19) and (20). K H and K H are factors describing the division of load between teeth and the load distribution on each tooth respectively. Generally K H can be set to 1. K H is more complicated since it only can be 1 in theory (if the gears are perfect). Here, for simplicity, it is set to 1.3, but if more exact data is available it should be used instead, see SS1871 for more information and guidelines about how to select this constant. Equation (27) gives the minimum size of the gear pair (with respect to Hertzian pressure) given a material, Hmax , E 1 , E 2 , 1 , 1 , a gear ratio, u, the number of pinion teeth, z 1 , the pressure angle and the calculated torque T cal . 3.2.2 Bending stress in the teeth roots The bending stress in a tooth, F can be calculated according to SS1871 as follows: bm KKF YYY FFcal FF = (28) The calculation of the form factor, Y F is somewhat complicated the way it is done in SS1871, therefore Y F is approximated with the following expression (Maskinelement handbok 2003): 14/ 1.32.2 z F eY + (29) Y F will always be larger for the small wheel (pinion) since it decreases with z. The helix angle factor Y is 1 for spur gears. Y is the so called contact ratio factor, and it is according to SS1871 calculated as follows: 1 =Y (30) Where the contact ratio, is calculated as before with equation (26). By combining equation (13), (23) and (28) the following is obtained: bua uzT KKYY cal FFFF 2 2 1 2 )1( + = (31) The expression above can be rewritten as follows: max 2 12 2 )1( F cal FFF u uzT KKYYba + = (32) Where Y F and Y is given by equation (29) and (30). K F and K F are factors describing the load division between teeth and the load distribution on each tooth respectively. If no other data is available K F can be set to 1, and K F to the same value as K H (SS1871). Equation (32) can be used to calculate the minimum size of a gear pair, with respect to bending endurance. Relations between size and gear ratio in spur and planetary gear trains 13(35) 3.2.3 Maximum allowed stress and pressure The maximum allowed bending stress Fmax and the maximum allowed hertzian pressure on the teeth, Hmax is of course largely dependent on material choice and safety factor. There are a lot of factors that can be included into the calculation of the stress limits, including the number of load cycles. Here the number of load cycles is assumed to be larger than the endurance limit and therefore that factor is disregarded. The maximum allowed stress and pressure is here simply calculated as follows: H H H S lim max = F F F S lim max = Where Hlim and Flim are material properties and S H and S F are the safety factors. For more advanced calculations see SS1871. Values of Flim and Hlim can for example be retrieved from Maskinelement Handbok (2003) or from the standards. It should be noted that if the safety factor for bending is doubled, the necessary gear size is doubled. But if the safety factor for the flank stress is doubled, it requires a four time larger gear pair, see equations (27) and (32) respectively. Relations between size and gear ratio in spur and planetary gear trains 14(35) 3.3 Results and sizing examples 3.3.1 Necessary Size/Volume In this section, equations (27) and (32) are applied on a (equivalent) load of 20 Nm. First, material data from an induction hardened steel with a Hertzian fatigue limit of 1200 MPa and a bending fatigue limit of 300 MPa is used. The Pressure angle is 20 degrees, and all other constants are set to the standard value (Table 1). Material Properties Gear Properties E 206 GPa 20 deg. 0.3 K H 1 7800 kg/m 3 K H 1.3 Load K F 1 T cal 20 Nm K F 1.3 Table 1. Values of material and gear parameters used in all examples. 0 5 10 15 10 15 20 25 30 35 40 0 0.5 1 1.5 2 2.5 3 3.5 4 x 10 5 Gear ratio Gear pair volumes (a 2 b). Hmax : 1200 MPa. Fmax : 300 Mpa T cal = 20 Nm No of teeth, wheel 1 (z 1 ) a 2 b m 3 0 5 10 15 10 15 20 25 30 35 40 0 0.2 0.4 0.6 0.8 1 1.2 x 10 4 Gear ratio Gear pair volumes (a 2 b). Hmax : 1200 MPa. Fmax : 300 Mpa T cal = 20 Nm No of teeth, wheel 1 (z 1 ) a 2 b m 3 root stress F1 root stress F2 flank stress H Figure 6. Gear pair volume as function of gear ratio and number of teeth. The safety factors SH and SF is 1 in the graph to the left and 2 in the graph to the right. The following plots are obtained for a non-hardened steel with a Hertzian fatigue limit of 500 MPa and a bending fatigue limit of 200 MPa (e.g. SIS1550). The load and all other parameters are the same as in the example above. 0 5 10 15 10 15 20 25 30 35 40 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 x 10 4 Gear ratio Gear pair volumes (a 2 b). Hmax : 500 MPa. Fmax : 200 Mpa T cal = 20 Nm No of teeth, wheel 1 (z 1 ) a 2 b m 3 0 5 10 15 10 15 20 25 30 35 40 0 1 2 3 4 5 6 7 8 x 10 4 Gear ratio Gear pair volumes (a 2 b). Hmax : 500 MPa. Fmax : 200 Mpa T cal = 20 Nm No of teeth, wheel 1 (z 1 ) a 2 b m 3 root stress F1 root stress F2 flank stress H Figure 7. The same plots as in Figure 6, but for another steel. The safety factors are set to one in the plot to the left and 2 in the plot to the right. Relations between size and gear ratio in spur and planetary gear trains 15(35) From the plots above and a number of other examples not presented here, a couple of conclusions can be made. First of all, the Hertzian pressure is the limiting factor in the vast number of cases (it requires the largest gears). The root bending stress is the limiting factor only for steels with a large difference between the two stress limits, and only if the safety factor for Hertzian pressure is low. Of course it is possible to change a lot of constants to get a different result, but these are the results if the standard (SS1871) is applied with the constants set to the recommended values. Furthermore the surface representing the root stress of the pinion (wheel 1) is always higher than corresponding surface for the gear wheel (wheel 2). This is consistent with the previous conclusion that the root stress always is larger for the smaller wheel. Therefore the root stress is only calculated for the pinion in the rest of this report. The number of teeth has most influence on the root bending stress. Not surprisingly, the flank stress is in comparison almost independent of the number of teeth. By choosing a relatively small number of pinion teeth (wheel 1), the tooth flank stress will almost certainly be the limiting factor. 3.3.2 Gear pair mass, geometry and inertia By combining equation (9) with the results shown in the left part of Figure 6 (safety factors = 1, induction hardened steel) the following graph (Figure 8) of the gear pair mass is obtained: 0 5 10 15 15 20 25 30 35 40 0 0.2 0.4 0.6 0.8 1 Gear ratio Gearpair mass as function of number of teeth and gear ratio, T = 20 No of teeth z 1 Mass kg Figure 8. Gear pair weight as function of number of teeth of the pinion and gear ratio. To continue the analysis it will be necessary to lock one of the variables in Figure 8, at least if the results are to be visualized in 3D-plots. The number of teeth of wheel 1 (pinion) seems to be the most reasonable variable to lock. In the figure below (Figure 9) plots of two different pinion teeth numbers are shown, 28 and 17. The larger choice results in, depending on the gear ratio, that booth the flank and root stress are limiting. The choice of 17 teeth of the small wheel results in, as also seen earlier, that only the flank stress is limiting the volume of the gear pair, regardless of gear ratio. Relations between size and gear ratio in spur and planetary gear trains 16(35) 0 5 10 15 0.5 1 1.5 2 2.5 3 3.5 x 10 5 Cross section when z1 = const = 28, T = 20 Nm a 2 b m 3 Gear Ratio Flank Stress Root Stress limiting (max) 0 5 10 15 0.5 1 1.5 2 2.5 3 x 10 5 Cross section when z1 = const = 17, T = 20 Nm a 2 b m 3 Gear Ratio Flank Stress Root Stress limiting (max) Figure 9. Cross sections of the left part of Figure 6, at 28 and 17 teeth of wheel one. To eliminate unrealistic gear geometry t