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Article

On Hirth Ring Couplings: Design Principles Including the Effect of Friction

1
Department of Industrial Engineering (DIN), University of Bologna, Viale del Risorgimento 2, 40136 Bologna, Italy
2
GIULIANI, Bucci Automations S.p.A. Division, Via Granarolo 167, 48018 Faenza, Italy
*
Author to whom correspondence should be addressed.
Actuators 2018, 7(4), 79; https://doi.org/10.3390/act7040079
Received: 21 October 2018 / Revised: 14 November 2018 / Accepted: 17 November 2018 / Published: 21 November 2018

Abstract

:
Rings with Hirth couplings are primarily used for the accurate positioning of axial-symmetric components in the machine tool industry and, generally, in mechanical components. It is also possible to use Hirth rings as connection tools. Specific industries with special milling and grinding machines are able to manufacture both tailor made and standard Hirth rings available on stock. Unfortunately, no international standard (for instance ISO, DIN or AGMA) is available for the production and the design of such components. In the best-case scenario, it is possible to find simplified design formulae in the catalogue of the suppliers. The aim of this work is to provide some accurate formulae and computational methods for design to provide better awareness on the limitations and the potential of this type of connection. The work consists of five parts: (i) a review of the base calculation derived mainly from the catalogues of manufacturers; (ii) an improved calculation based on a new analytical method including the friction phenomenon; (iii) an experimentation run for validating the method; (iv) a case study applied to a machine tool; and, (v) a closed form formulation to determine an upper threshold for friction, thus ensuring the Hirth coupling regular performance.

1. Introduction

Rings with Hirth couplings (Figure 1) are based on the concept of couplings achieved by frontal teeth connections.
The concept was invented and patented by a German engineer, Carl Albert Hirth in 1928 [1]. These couplings allow obtaining very high levels of:
accuracy, in terms of relative positioning between the parts connected together;
stiffness of the system, thanks to the steady self-center geometry of the teeth;
power, in terms of torque transmission;
resistance to wear, thanks to a wide support surface of the teeth and the use of special alloyed steels.
For these reasons, Hirth connections are successfully used in many mechanical applications: machine tools, turning tables, transfer units, turbo-chargers, robotics, ship building, and, recently, also in the automotive industry. With regard to the automotive industry, it is possible to highlight the example in Figure 2, dealing with the replacement of the standard splined shaft connection by a frontal teeth connection in the wheel hub. In conventional systems, the drive torque is transferred between the wheel bearing and the axle journal by means of radial splines, which are affected by a certain amount of clearance. Loads that occur in day-to-day driving conditions produce a risk of looseness beyond the normal mounting clearance, resulting in comfort reduction and unpleasant noises and, in some cases, the failure of the bearing. On the contrary, the self-centering axial gear teeth are placed on the axle journal and fixed by the central screw. The connection remains completely clearance-free in the gear teeth during the entire service life. Proceeding this way, it is possible to achieve up to 10% weight reduction per wheel (resulting in fuel consumption reduction and higher dynamic performance of the vehicle), a simplified assembly operation, a clearance free coupling and, above all, up to 150% of the torque transmission capability.
Hirth couplings, also regarded as curvic couplings, are widely used for torque transmission between mating discs in turbine machinery and are not covered by international standards (for instance ISO, DIN or AGMA). Their application makes it possible to fulfil an accurate centering, as well as an improved reliability and good structural stability. For instance, Reference [2] is focused on the structural and contact analysis of a curvic coupling in a gas turbine. In this study, a numerical model was developed to investigate the stress distribution and the contact evolution at different stages of the turbine service, in particular during preload, warm-up, speed-up, and running. The effects of coupling bolt preload, as well as of torsional load and of environmental temperature were also investigated. Other studies [3,4,5] have dealt with tooth contact aspects and with applications to turbo-engines in the aerospace industry [6,7].
A further interesting application of Hirth couplings, operating as servo-actuators, is in rotary tables for automatic machines and transfer machine tools [8]. These tables, fixed at predetermined angular positions, must be able to be fully constrained, thus resisting against high tangential forces. At the same time, they must have the capability of transmitting high torque, relying on tooth contact, rather than on friction couplings. Moreover, the Hirth coupling must ensure the accomplishment of strict specifications regarding the precise positioning of the rotating table, from the points of view of its angular position and especially its centering with respect to the machine axis, after rotation. The importance of correct alignment is also highlighted in [9,10], dealing with a similar issue.
Most papers in the field of transfer machine tools or of multiple-axis automatic machines are focused on modal or dynamic analyses, such as [11,12]. This research trend is also confirmed by a recent study [13] that deals with the development and validation of a multibody dynamic model of a tool-spindle-bearing system, to be incorporated into a five-axis machine with rotary-tilting spindle heads. Conversely, the topic of the stress peak at the tooth root has not been investigated so far. Moreover, no experimental validation has ever been conducted and published in the scientific literature on numerically computed stress peaks in Hirth couplings. A further unexplored point is the relationship between the stress peak and the actual response of the Hirth connection that may be affected by several factors, such as the force acting on each tooth, the contact area, or the friction affecting the device.
The subject of the present paper is to provide an original analytical model for the Hirth coupling design with regard to the peak stress estimation and to the presence of friction. As remarked above, issues of novelty arise from the lack of similar models in the literature, focused on the actual estimation of the stress maximum value for structural assessment. A further issue of novelty arises from the inclusion of the effect of friction (between the mating teeth or on the bearing surface) in the algorithm. In fact, this is usually disregarded in the standard procedure, but its effect deserves to be considered, as it may be even detrimental for safety and accuracy. The importance of taking friction into account is also emphasized in [2], where the developed curvic coupling finite element model accounts for this item. The present study, as a further evolution, proposes an experimentally validated closed form analytical approach that can be more efficiently applied for industrial design purposes and to properly arrange scheduled maintenance.

2. Standard Calculation

Matzke [14] provided some technical data able to give a rough overview of the calculation of Hirth rings. He mainly focused his investigation on the “nominal stress” in the tooth. Based on the nomenclature of Figure 3, the total tangential force Fu (N) is applied on the mean radius Rm (mm) of the G point (center of gravity) of the tooth section and can be evaluated by Equation (1) as a function of the external torque T (Nmm) to be transmitted:
F u = T R m  
Neglecting the effect of friction, for a tooth angle of 60° (in Hirth connection this angle is fixed to this value) the axial force generated by the coupling is given by Equation (2):
F a = F u · tan ( π 6 )  
The axial force can be produced, for example, by means of one or more bolts that must be accurately dimensioned [15,16,17,18,19]. The tooth is calculated against bending stress σb (MPa), approximating the trapezoidal section at the base by a rectangular one, as reported in Equation (3), where z is the number of teeth:
σ b = 6 · F u z · h G L · ( a + a 2 ) 2 σ b _ ref  
The reference values for the bending stress σb_ref (MPa) are reported in Table 1, based on the assumption of a minimum root radius (r > 0.3 mm [14]).
Technical German literature [20] provides a further formula in Equation (4) that is able to relate the torque T (Nmm) to the shear stress τ (MPa) to be compared to the allowable values τref (MPa) in Table 1.
τ = 16 · T π · D 3 [ 1 ( d D ) 4 ] τ ref  
The maximum (peak) bending stress at the tooth base depends on the stress distribution along the tooth and on the root radius r (mm) at the base. The values reported in [14,20] are able to provide nominal values only, without considering the local stress concentration. These simplified formulae have, anyway, the advantage of having been fully experimentally validated. In the most recent years, the development and improvements of finite element analysis have made it possible to achieve a more accurate estimation of the actual peak stress in mechanical components under defined hypothesis [21,22]. These methods can also be suitable for frontal teeth couplings [2,5,23]. The limitation in this field is related to the lack of experimental campaigns on real components, in order to relate the stress peak to the actual response of the Hirth connection. This is likely to be affected by several factors: the force acting on each tooth, the contact area, the holes for the installation, the friction between the teeth, the heat treatment process (case hardened teeth, not always with the same hardness), the hardness gradient on the teeth and the mechanical residual stresses due to the grinding operations.
Besides the analysis of Matzke, the technical literature on Hirth rings provides some more data for the designer, which are reported below. The calculation method is based on the assumption that the external torque T (Nmm) generates the tangential force Fu (N) (Equation (1)). Fu generates, in turn, the axial force Fa (N) (Equation (2)). This axial force Fa must be absorbed by preloading devices, such as disk springs, hydraulic pistons or bolts. The required pre-load Fv−a (N) is calculated, introducing a safety factor ν (Equation (5)):
F v a = F a · ν ν = 1.8     3
When compressed together, the teeth support each other if the preload Fva (N) is high enough. The varying loads on the tooth flanks result merely in a slightly irregular distribution of the pressure preload in the tooth root cross-section generating a maximum pressure pmax (MPa) according to Equation (6). In this equation (see Figure 4) Az (mm2) is the effective tooth flank area, D and d (mm) are the outer and the inner diameters of the teeth, dL (mm) is the fixing hole diameter, nb is the number of bolts in the teeth surface, r (mm) is the tooth root radius, s (mm) is the crown clearance, z the number of the teeth and ηz is the load bearing percentage (0.65 for milled teeth or 0.75 for grinded teeth). Under compression, with a high enough Fva load and with no transmission of torque, this load is equally distributed on both faces of each tooth. Conversely, when transmitting the torque T, the pressure rises on one face of the tooth and drops down on the other. The maximum pressure pmax (MPa) is calculated as follows:
{ p max = F v a + F a A z A z = ( D d n b · d L 2 D + d ) · [ π 4 ( D + d ) 1 . 155 · z · ( r + s ) ] · η z  
Concerning the geometry, the manufacturers offer a fixed parametric rule, in order to calculate the tooth height h (mm) in correspondence of the mean radius Rm (mm) (see Figure 4, Equation (7) and Table 2).
h = c · D ( 2 · r + s )  
Finally, concerning Hirth ring manufacturing, the most widely used materials are alloyed steels for quenching and tempering with Cr and Mo, namely 42CrMo4 (W. Nr. 1.7225) or 34Cr4 (W. Nr. 1.7033). Their production cycle is based on the following operations: turning, drilling the fixing holes and pins, milling the teeth, inductive hardening of the tooth area (hardening in the range 52 … 60 HRC), rough grinding of the teeth, finishing grinding of the teeth and of the reference surfaces, dimensional checking, finishing the pin holes, final checking and production of measuring report. An example of a measuring report is shown in Table 3.
The high precision and costs of these components allow an indexing accuracy of +/−2’’ (arcsec), a repeat accuracy <0.001 mm, a self-centering capability, a high wear resistance and a long-term service life.

3. Improved Calculation

3.1. Effect of Friction

The effect of friction between the mating teeth is not taken into account in the standard procedure. This effect is quite easy to be implemented and it is very important since it is detrimental to safety: according to Equation (8), in the presence of friction (angle of friction ρ (°)), with the same axial force, the actual transmission load Fu,μ (N) is reduced if the pre-load remains unchanged. Otherwise, a higher axial pre-load becomes necessary to achieve the same transmission. The coefficient of friction μ = tan(ρ) in compression couplings can be evaluated by experimental tests, such as those reported in [24,25], and can may range between 0.1 and 0.3 for smooth surfaces.
{ F u , μ = F a tan ( π 6 + ρ ) F u , μ = F u · K K = tan ( π 6 ) tan ( π 6 + ρ )  
In Figure 5 the reduction factor for the transmitted load (K, Equation (8)) is reported as a function of the angle of friction.
The pressure between the teeth is also affected by the coefficient of friction, but with a lower effect. By non-linear finite element analyses (non-linear due to sliding elements in contact), it is possible to calculate the percentage reduction factor for the tooth pressure (Kp) with respect to the value calculated by Equation (6)) and, then, to plot it in the diagram of Figure 6. In the case of negligible friction (μ = 0) the difference between Equation (6) and the numerical investigation is almost equal to zero.

3.2. Self-Centering Capability

It is well known that Hirth connections have self-centering capability. This feature is particularly important in all the applications in which the rings must be disengaged during their life. A very common example is the turntables of the machine tools. In a multi-station round transfer machine the sequence of two actions takes place: (i) the table is fixed with the Hirth rings engaged, when the machining operations are running (the rigid connection ensures the accomplishment of the strict geometrical tolerances of the machined parts); (ii) after the end of the machining tasks in one station, the table must rotate to the next station: for this purpose, the connections must be disengaged. Following the alignment with the next station, the rotary table must then be engaged and constrained again to let new machining operations start. Repeatability, accuracy and self-centering capability are necessary to produce parts compliant with very strict geometrical tolerances. Figure 7a explains the aforementioned cycle with regard to a triple Hirth connection: ring #1 is fixed in the machine bench and it is the reference for the accurate position (it does not move). Ring #2 is connected to the rotary table, so it moves (rotates) with the table (the table must return always in the same position with respect to the machine bench). Ring #3 is responsible for engaging/disengaging and for accurate position. When ring #3 moves up, the table is disengaged and can be rotated from station to station by a motor and a gearbox. Once the table gets close to the expected position, the motor is stopped in a pre-positioning area (as the allowance in the gears of the gearbox does not allow for a precise positioning), so ring #3 can move down. During its movement from the top to the bottom, since the teeth are conical from exterior to interior (Figure 3), ring #3 initially fits the reference ring #1 (so, ring #3 is aligned with ring #1). Afterwards, it fits the movable ring #2 that is therefore taken in the correct position (so, ring#2 is also aligned with ring #1). For the correct positioning of ring #2, the axial force Fa (N) has to be high enough to move all the weight related to the rotary table on the bearing surface. The calculation of the self-centering force Fc (N), which is not provided in the technical or in the scientific literature, will be shown in the following.
Three possible scenarios before the engaging phase have to be considered (see Figure 7b): (i) the fixed and the rotating ring have the same center and just have a different angular position; (ii) the fixed and the rotating ring have different centers and the same angular position; (iii) a combination of (i) and (ii).
Different configurations and alignments of teeth are highlighted in Figure 8: in case (a) the teeth of the fixed and rotating rings are engaged: they are all aligned and the space between the rings is constant (the upper ring is not represented); in case (b) before the engagement of the upper ring, there is an angular misalignment of the rotating ring. Since the rings have the same center, the space between the two rings is constant. In case (c) before the engagement of the upper ring, the two centers are not aligned (the self-centering force Fc should recreate the (a) condition). As a consequence, the space between the two rings is not constant on the 360° span and furthermore an angular misalignment γ (°) of the teeth also occurs: this is variable on the 360° span.
The conditions shown in Figure 8a,b (cases a and b) do not deserve any particular analytical evaluation. Conversely, the effect of the center misalignment c (mm) (Figure 8c, case c) requires a detailed analysis. According to Figure 9, the trend of the misalignment angle γ (°) around the 360° (θ between 0° and 360°) span is yielded by Equation (9):
{ r = R 2 + c 2 2 · R · c · cos ( θ ) γ = arcsin [ c r · sin ( θ ) ]  
As an example, for a ring diameter R = 450 mm, a center displacement c = 1 mm, when the angle θ is 29.5°, the angular misalignment γ is 0.063° (~4’); for θ = 89.5°, the angular misalignment γ is 0.127° (~8’). These results have also been checked by CAD system (Figure 9b). The whole trend is reported in Figure 9c for the interval 0–90°, provided that the function is symmetric every 90°.
In order to achieve the correct engagement, the following condition in Equation (10) must be fulfilled (the maximum misalignment γ must be within the angular step of the teeth):
γ max < 360 ° z  
Combining Equations (9) and (10), it is possible to work out the upper threshold cmax (mm) for the displacement of the center of the rotating ring, in order to ensure the correct position (Equation (11)). The maximum misalignment γ for θ = 90° (see Figure 9c) is considered for processing.
c max = r · sin ( 360 ° z ) R · sin ( 360 ° z )  
During the design phase, clearances and gaps in bearings, sealings and gears must accomplish this condition and the sum of all the assembly tolerances must be within cmax.
Unfortunately, the condition reported in Equation (11) is not sufficient to warrant the actual occurrence of self-centering; an additional force analysis and assessment is necessary. As a matter of fact, the self-centering force Fc (N) must be high enough to overcome all the friction forces, and to consequently move the rotating ring (along with the overall mass) towards its correct position. The self-centering force Fc (N) can be calculated by means of Equation (12), where every single tooth provides a different contribution, as a function of its angular position:
F c = 4 · i = 0 z / 4 F a z · tan ( π 6 + ρ ) · sin ( i · 360 ° z ) = K C · F a tan ( π 6 + ρ )  
with KC equal to 0.642 for z = 360; 0.650 for z = 144; 0.664 for z = 72.

4. Experimental Tests

Experimental tests on a triple Hirth ring connection with a mean diameter of 900 mm and z = 360 have been performed in order to check the proposed formulas (Figure 10).
At first, the static coefficient of friction μ in the bearing surface of the rotating ring (see Figure 7a) has been evaluated, in order to get its actual value. Thanks to a dynamometer, the radial force to be applied to move the rotating ring (merely supported on its bearing surface, without engagement) has been detected. The static coefficient of friction μ has therefore been calculated (five repeated tests) as the ratio between the radial force and its weight. The value of the static coefficient of friction μ is equal to 0.204 +/− 0.012; μ ~0.2 was consequently used for the calculation. Regarding the own weight of the upper ring (310 N) as the axial pre-loading force, Fa, and considering the angle of friction ρ between the teeth of 11.3° (μ = 0.2), a self-centering force Fc = ~225 N is yielded by Equation (12). The frictional resistance force Rμ (N) can be evaluated as the overall sum of the weights of the rotating and upper rings (i.e., the rings that must be moved into the correct position, leveraging the position of the fixed ring), multiplied by the aforementioned coefficient of friction in the bearing surface; a force Rμ = ~112 N is determined (reference sketch in Figure 11). The experimental test has shown the self-centering capability: the rings were able to shift to the correct engaged position, being Fc > Rμ. To fully validate the correctness of the formulation, an additional external force in the radial direction Fr,e (N) has been applied and measured by the same dynamometer (Figure 12). This force can be directly combined with the frictional force Rμ (it can be regarded as an extra-frictional force), because it is able to work against the self-centering force Fc. In this way, the self-centering capability has been evaluated accurately, step-by-step increasing the force Fr,e, until (Rμ + Fr,e) equalized the available Fc. The retrieved results are reported in Table 4; it can be highlighted that up to Fr,e = 100 N, the self-centering force is able to move the rings into the correct position, whereas in the range between Fr,e = 100 and 120 N, the self-centering capability is equal to the resistance forces: therefore, no movement occurs. As a matter of a fact, the calculated value of Fc = ~225 N (Equation (12)) is reliable since the addition of Rμ and Fr,e yields ~112 + 120 = 232 N.

5. Case Study

The case study proposed in this section refers to the rotary table of a big transfer machine tool, which has a Hirth ring connection with the same geometrical dimensions of that tested in the previous section. The present work moves from some issues related to the actual positioning of the rotary table (with concern regarding the loss of self-centering capability) after several years of work. After disassembling the rotary table, it was found that a lot of rust was present both in the teeth and in the bearing surface of the rotating ring (Figure 13). The presence of rust, as highlighted in [24,25], significantly increases the coefficient of friction, therefore reducing performance and self-centering capability. In this section, the aforementioned analytical predictive model is applied and the influence of the coefficient of friction on the actual performance is also studied.
The data for the calculation are reported in Table 5.
Analyzing the data in Table 5, based on force balance, it is possible to highlight that the following conditions are fulfilled in the presence of a coefficient of friction μ = ~0.2 (ID5).
  • The condition of the minimum torque T to be transmitted (ID8) is satisfied with the requested safety factor (SF = 4, ID11): 15,750 Nm > 3500 Nm (ID21 > ID8). The same condition also applies (obviously) to the tangential forces Fu: 35,000 N > 7778 N (ID13 > ID10).
  • The axial force Fa required for the transmission of the torque is 30,965 N (ID14): it is granted by a hydraulic piston (upon setting the pressure level of the oil, this force is fixed).
  • The frictional force Rμ, depending on the weight of the rotary table (40,000 N, ID16) and on the coefficient of friction (~0.2, ID5), is 8138 N. The self-centering capability is verified since the condition Fc > Rμ is satisfied (22,326 N > 8138 N, ID15 > ID17).
  • The accuracy of the angular positioning is also verified. The friction torque on the bearing surface of the rotating ring, which depends on the frictional force Rμ, on the mean sliding radius (ID18) and on the coefficient of friction (ID5), is 3825 Nm; the additional frictional torque for the rotating elements (such as rotary sealings for oil distribution, bearings and gears) is 3000 Nm (ID20). The torque T that the Hirth connection is able to provide proves to be higher than the total frictional torque Tμ: 15,750 Nm > 6825 Nm (ID21 > ID22).
Concerning the stress analysis, it is easy to run this by Equation (3) or (4), with the teeth geometry reported in Figure 14 and considering the full torque capability T = 15,750 Nm. The equations above yield: σb = 0.23 MPa and τ = 0.60 MPa, which are very far from their admissible thresholds (50 MPa and 18.5 MPa respectively, Table 1).
Finally, the effect of the coefficient of friction is shown. According to [24,25], the coefficient of friction is strongly affected by the presence of the rust, which makes it increase. For this reason, a sensitivity analysis has been performed. The axial force Fa has not been modified, provided that the system works with a hydraulic piston set at a fixed pressure level and is able to provide Fa = 30,965 N (ID14). Fu,μ is therefore recalculated by Equation (8), depending on the actual value of the friction angle ρ. The other forces depending on friction are also re-calculated. The results are reported in the chart of Figure 15 and in Table 6, where the double effect of the increase of the coefficient of friction is well visible; on one hand, it reduces the capability of the active (self-centering) forces and on the other, it increases the resistance (frictional) forces. In the proposed case study, the detrimental effect of rust is very clear: an increase of the coefficient of friction from its initial value of ~0.2 (for new surfaces in contact) to a value of 0.4 (or more for the rusted ones) results in the complete loss of the positioning capability of the Hirth connection for the rotary table. This in turn implies the loss of accuracy and precision in the production of the components manufactured by the transfer machine tool. When this situation occurs, the machine tool requires maintenance.

6. Closed-Form Determination of a Friction Threshold to Ensure Self-Centering

The outcome of the previous analysis suggests the existence of an upper threshold for friction. For a fixed Hirth geometry, this depends on the applied axial force, Fa, and on the actual overall weight to be moved, W. When friction at the interface is lower than this threshold, the centering force is able to warrant full alignment. Conversely, when friction is above this level, due to wear or rust, the axial force is no longer sufficient to allow centering.
Considering Equation (12), the friction threshold can be determined by imposing that the centering force, Fc, is greater than the frictional force arising from the weight to be moved, Rμ. This condition is highlighted in Equation (13), where μ indicates the frictional coefficient and ρ the corresponding frictional angle:
F c = K c · F a tan ( π 6 + ρ ) R μ = μ · W = tg ( ρ ) · W  
Some simple algebraic operations yield the solution in terms of the upper threshold for the frictional coefficient, μth:
μ th . = 3 3 ( K c F a + W ) + 1 3 ( K c F a + W ) 2 + 4 WK c F a 2 W  
This threshold is plotted in Figure 16, as a function of the applied axial force, for different overall weights to be moved for KC = 0.642 (z = 360). It can be observed that μth. exhibits a non-linear increasing trend. The higher the axial force that can be transmitted by the third ring, the higher the frictional threshold, which even tends to unity. Remarkably high values of μth. clearly indicate that the frictional force component can be always easily overcome by the centering force, regardless of the presence of rust or wear. In addition, the higher the weight to be moved, the lower the retrieved threshold. This outcome indicates that, when large masses have to be centered, remarkably high axial forces are required to overcome friction, even for a low frictional coefficient at the interface. Finally, it is interesting to observe that the results in the diagram below are consistent with those of the case study above. In fact, for an axial force of around 31 kN and a weight of 40 kN, the friction threshold is 0.395. This result confirms the unverified condition for centering, as μ increases up to 0.4, as shown in Table 6.

7. Conclusions

This paper presents a methodology for the precise calculation of the forces generated in Hirth ring connections. The role of friction is taken into account and is demonstrated to be fundamental for the correct calculation of this type of component. Simplified formulae in scientific or technical literature are not sufficiently accurate to describe the performance of Hirth rings in full detail. Starting from a practical example that occurred in a working machine tool that lost the accuracy of its positioning capability, some more advanced design formulae have been proposed. These novel equations have also been experimentally verified. Friction conditions are likely to change during the lifespan of a machine tool, mainly due to wear, rust and corrosion. Therefore, the designer needs to use a reliable design method for Hirth connection calculations for long-term life. The developed model, which incorporates the effect of friction, is very important to predict the capability of the connection, and also indicates the related friction thresholds to achieve self-centering and adequate torque transfer.

Author Contributions

Conceptualization, D.C. and N.V.; Data curation, M.D.A., S.F., F.R. and N.V.; Formal analysis, G.O. and N.V.; Investigation, M.D.A., S.F. and N.V.; Methodology, D.C., M.D.A., S.F., G.O. and N.V.; Project administration, D.C. and N.V.; Supervision, D.C., G.O. and N.V.; Validation, N.V.; Writing—original draft, G.O. and N.V.; Writing—revised version, G.O., F.R. and N.V.

Conflicts of Interest

The authors declare no conflict of interest.

Nomenclature

AzEffective tooth flank area (mm2)
a’, a”, hGGeometrical features of the Hirth ring and related teeth (from Figure 3)
cCenter misalignment between the fixed and the rotating rings (mm)
cmaxMaximum center misalignment between the fixed and the rotating rings (mm)
DOuter diameter of the teeth (mm)
dInner diameter of the teeth (mm)
dLFixing hole diameter (mm)
Factor cFactor depending on the number of teeth z (from Table 2)
FaAxial force generated by the Hirth coupling (N, kN)
FcSelf-centering force (N, kN)
Fr,eAdditional external force in the radial direction (experimental validation test) (N)
FSSafety factor for the total tangential force (-)
FuTotal tangential force (N, kN)
Fu,μActual tangential force (including friction) (N, kN)
Fν-aPreload (N, kN), considering a safety factor ν
hTooth height (mm)
KReduction factor for the transmitted load (-)
KCCoefficient for the self-centering force computation, depending on the number of teeth
KpPercentage reduction factor for the tooth pressure (%)
nbNumber of bolts in the teeth surface
pmaxMaximum pressure on tooth flank (MPa)
RRing diameter (mm)
RmMean radius of the ring (mm)
RμFrictional resistance force (N, kN)
rTooth base radius (mm)
sCrown clearance (mm)
TExternal torque to be transmitted (Nm, Nmm)
TµFrictional torque (Nm, Nmm)
WOverall weight to be moved upon self-centering (N, kN)
zNumber of teeth (-)
γAngular misalignment between the fixed and the rotating rings (°)
γmaxMaximum angular misalignment between the fixed and the rotating rings (°)
ηzLoad bearing percentage (-)
μCoefficient of friction (-)
μthCoefficient of friction upper threshold to achieve self-centering (-)
νSafety factor for preload (-)
θGeneral angular coordinate over the Hirth ring (°)
ρAngle of friction (°, rad)
σbTooth bending normal stress (MPa)
σb_refAllowable normal stress (MPa)
τShear stress due to the transmitted torque (MPa)
τ_refAllowable shear stress (MPa)

References

  1. Hirth, C.A. Shaft Coupling. U.S. Patent 1,660,792, 28 February 1928. [Google Scholar]
  2. Yuan, S.X.; Zhang, Y.Y.; Zhang, Y.C.; Jiang, X.J. Stress distribution and contact status analysis of a bolted rotor with curvic couplings. Proc. I. Mech. Eng. C-J. Mec. 2010, 224, 1815–1829. [Google Scholar] [CrossRef]
  3. Pisani, S.R.; Rencis, J.J. Investigating CURVIC coupling behavior by utilizing two- and three-dimensional boundary and finite element methods. Eng. Anal. Bound. Elem. 2000, 24, 271–275. [Google Scholar] [CrossRef]
  4. Richardson, I.J.; Hyde, T.H.; Becker, A.A.; Taylor, J.W. A validation of the three-dimensional finite element contact method for use with curvic couplings. Proc. Inst. Mech. Eng. G-J. Aerosp. Eng. 2002, 216, 63–275. [Google Scholar] [CrossRef]
  5. Richardson, I.J.; Hyde, T.M.; Becker, A.A.; Taylor, J.W. A three-dimensional finite element investigation of the bolt stresses in an aero-engine curvic coupling under a blade release condition. Proc. Inst. Mech. Eng. G-J. Aerosp. Eng. 2000, 214, 231–245. [Google Scholar] [CrossRef]
  6. Yin, Z.Y.; Hu, B.A.; Wu, J.G.; Xu, Y.L.; Zheng, Q.X. Calculation of axial relaxed/pressed forces of rotors with curvic couplings. Acta Aeronaut. Aeronaut. Sin. 1996, 17, 555–560. [Google Scholar]
  7. Hu, B.A.; Yin, Z.Y.; Xu, Y.L. Determination of axial preloads of rotor with curvic couplings pretightened into two segments. J. Mech. Strength 1999, 21, 274–277. [Google Scholar]
  8. Croccolo, D.; Cavalli, O.; De Agostinis, M.; Fini, S.; Olmi, G.; Robusto, F.; Vincenzi, N. A Methodology for the Lightweight Design of Modern Transfer Machine Tools. Machines 2018, 6, 2. [Google Scholar] [CrossRef]
  9. Cao, H.; Li, D.; Yue, Y. Root Cause Identification of Machining Error Based on Statistical Process Control and Fault Diagnosis of Machine Tools. Machines 2017, 5, 20. [Google Scholar][Green Version]
  10. Zhang, F.P.; Lu, J.P.; Tang, S.Y.; Sun, H.F.; Jiao, L. Locating error considering dimensional errors modeling for multistation manufacturing system. Chin. J. Mech. Eng. 2010, 23, 765–773. [Google Scholar] [CrossRef]
  11. Du, C.; Zhang, J.; Lu, D.; Zhang, H.; Zhao, W. A parametric modeling method for the pose-dependent dynamics of bi-rotary milling head. Proc. Inst. Mech. Eng. B-J. Eng. Manuf. 2018, 232, 797–815. [Google Scholar] [CrossRef]
  12. Liu, X.; Yuan, Q.; Liu, Y.; Gao, J. Analysis of the stiffness of hirth couplings in rod-fastened rotors based on experimental modal parameter identification. In Proceedings of the ASME Turbine Technical Conference and Exposition, Düsseldorf, Germany, 16–20 June 2014. [Google Scholar]
  13. Du, C.; Zhang, J.; Lu, D.; Zhang, H.; Zhao, W. Coupled Model of Rotary-Tilting Spindle Head for Pose-Dependent Prediction of Dynamics. J. Manuf. Sci. Eng. 2018, 140, 081008. [Google Scholar] [CrossRef]
  14. Matzke, G. Verbindung von Wellen durch Verzahnung. Konstruktion 1951, 3, 211–216. [Google Scholar]
  15. Croccolo, D.; De Agostinis, M.; Fini, S.; Olmi, G. Tribological properties of bolts depending on different screw coatings and lubrications: An experimental study. Tribol. Int. 2017, 107, 199–205. [Google Scholar] [CrossRef]
  16. Croccolo, D.; De Agostinis, M.; Vincenzi, N. A contribution to the selection and calculation of screws in high duty bolted joints. Int. J. Pres. Ves. Pip. 2012, 96, 38–48. [Google Scholar] [CrossRef]
  17. Croccolo, D.; De Agostinis, M.; Vincenzi, N. Influence of tightening procedures and lubrication conditions on titanium screw joints for lightweight applications. Tribol. Int. 2012, 55, 68–76. [Google Scholar] [CrossRef]
  18. Croccolo, D.; De Agostinis, M.; Fini, S.; Olmi, G. An experimental study on the response of a threadlocker, involving different materials, screw dimensions and thread proportioning. Int. J. Adhes. 2018, 83, 116–122. [Google Scholar] [CrossRef]
  19. Croccolo, D.; Vincenzi, N. Tightening tests and friction coefficients definition in the steering shaft of front motorbike suspension. Strain 2011, 47, 337–342. [Google Scholar] [CrossRef]
  20. Niemann, G.; Winter, H.; Hohn, B.R. Maschinenelemente; Springer-Verlag: Berlin, Germany, 2005. [Google Scholar]
  21. Croccolo, D.; De Agostinis, M.; Vincenzi, N. Normalization of the stress concentrations at the rounded edges of a shaft-hub interference fit: Extension to the case of a hollow shaft. J. Strain Anal. Eng. 2012, 47, 131–139. [Google Scholar] [CrossRef]
  22. Croccolo, D.; De Agostinis, M.; Fini, S.; Morri, A.; Olmi, G. Analysis of the influence of fretting on the fatigue life of interference fitted joints. In Proceedings of the ASME International Mechanical Engineering Congress and Exposition, Montreal, QC, Canada, 14–20 November 2014; Volume 2B, p. V02BT02A008. [Google Scholar] [CrossRef]
  23. Jiang, X.J.; Zhang, Y.Y.; Yuan, S.X. Analysis of the contact stresses in curvic couplings of gas turbine in a blade-off event. Strength Mater. 2012, 44, 539–550. [Google Scholar] [CrossRef]
  24. Croccolo, D.; Cuppini, R.; Vincenzi, N. Friction coefficient definition in compression-fit couplings applying the DOE method. Strain 2008, 44, 170–179. [Google Scholar] [CrossRef]
  25. RR71, Friction in Temporary Works, ISBN 0 7176 2613 X, HSE Books. Available online: http://www.hse.gov.uk/research/rrhtm/rr071.htm (accessed on 19 November 2018).
Figure 1. Example of Hirth rings: (a) disengaged and (b) engaged.
Figure 1. Example of Hirth rings: (a) disengaged and (b) engaged.
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Figure 2. Example of Hirth connection in automotive: (a) old design with radial teeth and (b,c) new design with frontal teeth.
Figure 2. Example of Hirth connection in automotive: (a) old design with radial teeth and (b,c) new design with frontal teeth.
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Figure 3. Nomenclature and forces acting on the tooth.
Figure 3. Nomenclature and forces acting on the tooth.
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Figure 4. Geometric parameters from the manufacturer “Voith-Turbo Gmbh”.
Figure 4. Geometric parameters from the manufacturer “Voith-Turbo Gmbh”.
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Figure 5. Reduction factor of transmission load K as a function of the angle of friction ρ.
Figure 5. Reduction factor of transmission load K as a function of the angle of friction ρ.
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Figure 6. Reduction factor for the tooth pressure Kp as a function of the coefficient of friction µ.
Figure 6. Reduction factor for the tooth pressure Kp as a function of the coefficient of friction µ.
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Figure 7. (a) Example of rotary table for the transfer machine tool; (b) position of ring #2 before engaging phase: left—the angle has to be adjusted; right—the center has to be adjusted.
Figure 7. (a) Example of rotary table for the transfer machine tool; (b) position of ring #2 before engaging phase: left—the angle has to be adjusted; right—the center has to be adjusted.
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Figure 8. (a) Ring #1 and ring #2 engaged correctly; (b) before engagement: pure angular misalignment; (c) before engagement: center misalignment (which also generates angular misalignment).
Figure 8. (a) Ring #1 and ring #2 engaged correctly; (b) before engagement: pure angular misalignment; (c) before engagement: center misalignment (which also generates angular misalignment).
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Figure 9. (a) Sketch for the evaluation of the misalignment angle γ; (b) practical example via CAD system and (c) related trend of γ as a function of θ.
Figure 9. (a) Sketch for the evaluation of the misalignment angle γ; (b) practical example via CAD system and (c) related trend of γ as a function of θ.
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Figure 10. Hirth ring connection with a mean diameter of 900 mm and z = 360.
Figure 10. Hirth ring connection with a mean diameter of 900 mm and z = 360.
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Figure 11. Hirth self-centering action.
Figure 11. Hirth self-centering action.
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Figure 12. Self-centering capability with the presence of Rμ and Fr,e.
Figure 12. Self-centering capability with the presence of Rμ and Fr,e.
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Figure 13. Comparison between the status of a new Hirth connection vs. a rusty one.
Figure 13. Comparison between the status of a new Hirth connection vs. a rusty one.
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Figure 14. Tooth geometry for the case study.
Figure 14. Tooth geometry for the case study.
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Figure 15. Sensitivity analysis with the coefficient of friction for (a) forces and (b) torque.
Figure 15. Sensitivity analysis with the coefficient of friction for (a) forces and (b) torque.
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Figure 16. Friction threshold and its dependence on the axial force and the weight to be moved.
Figure 16. Friction threshold and its dependence on the axial force and the weight to be moved.
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Table 1. Allowable stresses (σb_ref and τ_ref) for carbon steel and for alloyed steels for base radius r > 0.3 mm [14,20].
Table 1. Allowable stresses (σb_ref and τ_ref) for carbon steel and for alloyed steels for base radius r > 0.3 mm [14,20].
Type of StressCarbon SteelAlloyed Steel
Cr-Ni and Cr-Mo
σb_ref (MPa)τ_ref (MPa)σb_ref (MPa)τ_ref (MPa)
No Shocks9033.512044.5
With shocks5018.57026
With shocks and with torsional vibration35135018.5
Table 2. Tables provided by the manufacturer “Voith-Turbo Gmbh” for the proper choice of Factor c and of the crown clearance s.
Table 2. Tables provided by the manufacturer “Voith-Turbo Gmbh” for the proper choice of Factor c and of the crown clearance s.
Number of Teeth zFactor c
120.234
240.114
360.075
480.056
600.045
720.037
960.028
1200.022
1440.018
1800.015
2400.011
2880.009
3600.007
7200.003
Tooth Root Radius rCrown Clearance s
0.30.4
0.60.6
1.01.0
1.61.6
2.52.5
Table 3. Measuring report of Hirth rings engagements.
Table 3. Measuring report of Hirth rings engagements.
ControlTheoretical MeasuresToleranceResults
∅ ext.C1
C2990 mm−0.01 to −0.03 mm989.971 mm
C3
∅ ext.
Circularity
C1
C2 0.005 mm
C3
∅ int.C1849 mm H7+0.090 mm849.018 mm
C2
C3829 mm+0.030 mm829.017 mm
∅ int.
Circularity
C1 0.004 mm
C2
C3 0.004 mm
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Concentricity
∅ int.- ∅ ext.
C1 Actuators 07 00079 i002 0.003 mm
C2 Actuators 07 00079 i003 0.004 mm
C3 Actuators 07 00079 i004 0.003 mm
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Flatness
C1 0.003 mm
C2 0.002 mm
C3 0.002 mm
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Parallelism
C1-C2
Rotating C2
Actuators 07 00079 i007 0.007 mm
90° 0.009 mm
180° 0.008 mm
270° 0.009 mm
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Parallelism
C1-C3
Rotating C3
Actuators 07 00079 i009 0.008 mm
90° 0.009 mm
180° 0.006 mm
270° 0.009 mm
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Concentricity
C1-C2
Rotating C2
Actuators 07 00079 i011 0.008 mm
90° 0.009 mm
180° 0.006 mm
270° 0.009 mm
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Concentricity
C1-C3
Rotating C3
Actuators 07 00079 i013 0.007 mm
90° 0.009 mm
180° 0.008 mm
270° 0.009 mm
Indexing accuracy 1.50’’
HardnessHRC 54±2 (HRC)HRC 54
Table 4. Experimental verification of the self-centering capability.
Table 4. Experimental verification of the self-centering capability.
Additional External Radial Load Fr,e (N)Self-Centering Capability (YES/NO)
20YES Actuators 07 00079 i014
40YES Actuators 07 00079 i014
60YES Actuators 07 00079 i014
80YES Actuators 07 00079 i014
100YES Actuators 07 00079 i015
120NO Actuators 07 00079 i016
Table 5. Calculation data for a rotary table of a transfer machine (Figure 7a).
Table 5. Calculation data for a rotary table of a transfer machine (Figure 7a).
IDDescriptionType of DataValueUnits
1Tooth AngleInput data (fixed for Hirth connection)30°
2 0.524rad
3Friction angle Input data11.5°
4 0.201rad
5Coefficient of friction µCalculated: µ = tan(ρ)0.203
6Maximum external tangential force to be transmitted during a milling operationInput data5000N
7Distance from the center of the tangential force (radius of the rotary table)Input data700mm
8Minimum Torque T to be transmittedCalculated: ID6 × ID73500Nm
9Hirth rings mean radiusInput data450mm
10Tangential force Fu on the Hirth connectionCalculated: ID8 × 1000/ID97778N
11Safety factor SFInput data (reference value: 3 … 5)4
12Tangential force Fu on the Hirth connection including SFCalculated: ID10 × ID1131,111N
13Tangential force Fu used for the designSelected by the designer on the basis of ID1235,000N
14Axial force Fa (including friction)Calculated (info in the paper)30,965N
15Self-centering capability FcCalculated (info in the paper)22,326N
16Weight of the rotary table connected to the rotary ringInput data40,000N
17Frictional force RµCalculated: ID16 × ID58138N
18Mean sliding radius of the bearing surfaceInput data470mm
19Frictional torque on the bearing surfaceCalculated: ID17 × ID183825Nm
20Additional frictional torque (sealings, bearings, …)Input data3000Nm
21Torque T that Hirth rings can transmitCalculated: ID13 × ID915,750Nm
22Frictional torque TµCalculated: ID19 + ID206825Nm
Table 6. Sensitivity analysis for different values of the friction coefficient.
Table 6. Sensitivity analysis for different values of the friction coefficient.
Active Actions Provided by the HirthResistance Actions Due to FrictionCheck
Coefficient of Friciton μAngle of Friction ρ (°)Axial Force Fa from Hydraulic Piston (N)Tangential Force Fu (N)Self-Centering Force Fc (N)Torque Available for Correct Angular Alignment (Nm)Frictional Force Rμ (N)Frictional Torque Tμ (Nm)Fc > RμT > Tμ
0.20011.330,96535,23422,43115,85580006760YESYES
0.30016.729,18118,57713,13112,0008640YESYES
0.40021.824,36615,51210,96516,00010,520NOYES
0.50026.620,44513,016920020,00012,400NONO
0.60031.017,19010,944773524,00014,280NONO

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MDPI and ACS Style

Croccolo, D.; De Agostinis, M.; Fini, S.; Olmi, G.; Robusto, F.; Vincenzi, N. On Hirth Ring Couplings: Design Principles Including the Effect of Friction. Actuators 2018, 7, 79. https://doi.org/10.3390/act7040079

AMA Style

Croccolo D, De Agostinis M, Fini S, Olmi G, Robusto F, Vincenzi N. On Hirth Ring Couplings: Design Principles Including the Effect of Friction. Actuators. 2018; 7(4):79. https://doi.org/10.3390/act7040079

Chicago/Turabian Style

Croccolo, Dario, Massimiliano De Agostinis, Stefano Fini, Giorgio Olmi, Francesco Robusto, and Nicolò Vincenzi. 2018. "On Hirth Ring Couplings: Design Principles Including the Effect of Friction" Actuators 7, no. 4: 79. https://doi.org/10.3390/act7040079

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