CN116465550A - Planetary gear unbalance force determination method, evaluation method and system comprising multiple non-fixed rotators - Google Patents
Planetary gear unbalance force determination method, evaluation method and system comprising multiple non-fixed rotators Download PDFInfo
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Abstract
The invention belongs to the technical field of planetary speed change mechanisms, and particularly discloses a method for determining a planetary gear set dynamic unbalance force comprising a plurality of non-fixedly connected revolution bodies, an evaluation method and a system, wherein the determination method is used for calculating the eccentricity of a clutch oil cylinder equivalent turntable and the total eccentricity of the clutch equivalent turntable based on basic parameters of a clutch in a non-fixedly connected revolution body component, and then calculating the eccentricity of a planetary gear set frame equivalent turntable; calculating the total eccentricity of the planet row frames according to the determined eccentricity of each planet row frame caused by the positioning error, the mass deviation, the tooth side gap between the inner gear ring and the planet wheel and the support error of the frames; and calculating dynamic unbalance force and dynamic unbalance moment corresponding to each equivalent turntable of the planetary speed change mechanism according to the total eccentricity of the clutch and the total eccentricity of the planetary row frame. By adopting the technical scheme, the eccentric quantity and the deflection quantity of each revolving body of the speed change mechanism under various influencing factors are fully considered, and accurate dynamic unbalance force evaluation is realized.
Description
Technical Field
The invention belongs to the technical field of planetary speed change mechanisms, and relates to a method for determining and evaluating a planetary gear movement unbalance force comprising a plurality of non-fixedly connected revolution bodies and a system thereof.
Background
The factors such as uneven material quality of the revolving body, assembly error of the revolving body, deformation and displacement in the running process of the revolving body can cause misalignment of the mass center (inertia axis) and the rotation center (geometric axis) of the revolving body, and finally, the revolving body can generate additional dynamic unbalance force when rotating at high speed. Dynamic unbalance forces are important excitation sources on planetary transmission (planetary row) rotor systems. Because the parts such as the planet row frame, the gears and the operating parts in the planetary speed change mechanism belong to non-fixedly connected revolution body systems, and gaps are necessarily formed in all revolution bodies, the revolution body systems are necessarily eccentric and inclined, and dynamic unbalance force is generated during high-speed operation. The dynamic unbalance force can cause excessive vibration and stress when the system rotates at a high speed, accelerate the abrasion of parts such as gears, bearings and the like, deteriorate the working environment of the planet row and cause energy waste, and simultaneously limit the highest working rotating speed of the planet row.
The dynamic unbalance force of a non-fixed revolving body in a planetary speed change mechanism is influenced by a plurality of factors, and an effective and reliable dynamic unbalance force determining and evaluating method is not formed for a planetary transmission system comprising a plurality of non-fixed revolving bodies by using the existing dynamic unbalance force calculating method only aiming at a single revolving body.
Disclosure of Invention
The invention aims to provide a method, a method and a system for determining the dynamic unbalance force of a planetary gear train with multiple non-fixed rotators, and the method and the system for evaluating the dynamic unbalance force accurately.
In order to achieve the above purpose, the basic scheme of the invention is as follows: a method for determining the unbalanced force of planetary gear set with multiple non-fixed rotators includes the following steps:
s1, acquiring basic parameters of a non-fixedly connected revolving body assembly of a planetary gear transmission mechanism to be calculated;
s2, calculating the eccentricity of an equivalent turntable of the clutch oil cylinder caused by the eccentricity of the clutch oil cylinder, the eccentricity of a friction plate and a spline clearance based on basic parameters of the clutch in the non-fixedly connected revolving body assembly;
s3, calculating the total eccentricity of the clutch equivalent turntable according to the eccentricity of the clutch equivalent turntable caused by the eccentricity of the oil cylinder, the eccentricity of the friction plate and the spline clearance determined in the S2;
s4, calculating the eccentricity of an equivalent turntable of the planet row frame, which is caused by the positioning error, the mass deviation, the tooth side clearance between the inner gear ring and the planet wheel, and the support error of the frame, of the planet row frame in the non-fixedly connected revolving body component based on basic parameters of the planet row frame;
s5, calculating the total eccentricity of the planet row frames according to the eccentricity of each planet row frame caused by the positioning error, the mass deviation, the tooth side gap between the inner gear ring and the planet wheel and the support error of the frames, which are determined in the step S4;
and S6, calculating dynamic unbalance force and dynamic unbalance moment corresponding to each equivalent turntable of the planetary speed change mechanism according to the total eccentricity of the clutch calculated in the step S3 and the total eccentricity of the planetary row frame calculated in the step S5.
The working principle and the beneficial effects of the basic scheme are as follows: according to the technical scheme, the eccentric quantity and the deflection quantity of each revolving body of the speed change mechanism under various influencing factors are fully considered, the higher calculation efficiency is ensured, meanwhile, the situation is closer to the actual situation, and more accurate dynamic unbalance force and dynamic unbalance moment are obtained and used for measuring the high-speed running performance of the speed change mechanism.
Further, the non-fixedly connected revolving body component comprises a clutch, a planet row frame, a planet wheel and an inner gear ring;
the basic parameters of the non-fixedly connected revolving body component comprise the number of teeth, the modulus and the reference circle pressure angle of a friction plate and an inner hub in a clutch, the mass of the friction plate and the inner hub, the angular velocity of the friction plate, the mass and the angular velocity of the clutch, the mass and the angular velocity of a planet row frame, the mass and the angular velocity of an inner gear ring, the center distance between a planet wheel and a sun wheel, the center distance between the inner gear ring and the planet wheel, the diameter of an inner ring and the outer ring of a support bearing and the tolerance level.
The required parameters are obtained, and the subsequent use is facilitated.
Further, the method for calculating the eccentricity of the equivalent turntable of the clutch oil cylinder caused by the eccentricity of the clutch oil cylinder, the eccentricity of the friction plate and the spline clearance in the step S2 is as follows:
according to basic parameters of clutch cylinder and the eccentric value e of cylinder obtained by experimental measurement 0 The eccentricity equivalent to the clutch turntable is e d1 ：
Wherein m is 0 The mass of the oil cylinder; omega 0 Is the angular velocity of the oil cylinder; m is M d The mass of the clutch turntable; omega d The angular speed of the clutch is the same as the rotational speed of the input shaft;
according to basic parameters of the inner hub and the friction plate, estimating the relation between the eccentric displacement and the spline tooth side gap c by utilizing geometric kinematics, and calculating the initial eccentricity a caused by the tooth side gap on the assumption that the single-side tooth side gap is subjected to Gaussian random distribution:
c i ＝normrnd(μ,σ)
a＝min(a left ,a right )
wherein c i A single-sided tooth flank clearance for the ith tooth; norm represents the generation of random numbers subject to gaussian distribution; μ is the average value of the tooth flank clearance; sigma (sigma)For the degree of dispersion of the backlash, the backlash ranges between (μ -3σ, μ+3σ); a, a left The eccentricity of the contact of the left side of the tooth before the right side of the tooth; a, a right The eccentricity of the contact of the right side of the tooth before the left side of the tooth; abs represents absolute value; n is the number of teeth; alpha is the pressure angle;
by analyzing the eccentricity caused by the backlash under different numbers of teeth, the eccentricity is obtained to be in a decreasing trend along with the increase of the number of teeth, and the larger the number of teeth is, the closer the number of teeth is to the backlash, so when the elastic deformation of the friction plate or the spline teeth is not considered, the eccentricity of the friction plate caused by the clearance c is considered to be the same as the clearance, namely e f ＝c；
Based on the derived eccentricity e of the friction plate f Equivalent to the eccentricity e of the clutch disc d2 The method comprises the following steps:
e d2 ＝Nm f e f ω f 2 /M d ω d 2
wherein N is the number of friction plates of the clutch; m is m f Is the mass of the friction plate; omega f Is the angular velocity of the friction plate; omega d The angular speed of the clutch is the same as the rotational speed of the input shaft;
determining the eccentricity e of the spline side gap on the clutch turntable according to the involute spline clearance fit category and the spline tooth tolerance grade in national standard GB/T3748.1-2008 d3 。
And the corresponding eccentricity of the clutch is calculated, so that the clutch is convenient to use.
Further, the method for calculating the total eccentricity of the equivalent rotary disc of the clutch in the step S3 is as follows:
total eccentricity e of clutch dsum Calculating the resulting eccentricity e equivalent to the clutch disc for step S2 d1 Eccentric e equivalent to clutch rotor disk d2 And eccentricity e of the spline backlash to the clutch disc d3 If the eccentric direction caused by each factor is the same direction, the maximum eccentric amount of the clutch is calculated as (e dsum ) max The method comprises the following steps:
(e dsum ) max ＝e d1 +e d2 +e d3
calculate the comprehensive eccentric amount (e) dsum ) sys The method comprises the following steps:
the operation is simple, and the operation is facilitated.
Further, in step S4, the method for calculating the eccentricity of the equivalent turntable of the planet carrier caused by the positioning error, mass deviation, backlash between the ring gear and the planet carrier, and support error of the carrier is as follows:
according to the mass m of the planet wheel 1 And planet wheel position error delta e, calculating equivalent eccentric quantity e of planet row frame p1 The method comprises the following steps:
wherein omega 1 Is the angular velocity of the planet wheel; m is M p Equivalent mass for the planet carrier; omega p Angular velocity for the planet row frame;
calculating the equivalent eccentric quantity e of the planet row frame according to the mass deviation delta m of the planet wheel of the speed change mechanism p2 The method comprises the following steps:
wherein r represents the distance from the center of the planet wheel hole to the rotation center of the planet row frame;
selecting a gear side gap between an inner gear ring and a planet gear according to the center distance of a gear pair, and decentering e caused by the gear side gap between the inner gear ring and the planet gear in a planet gear frame pr About equal to the tooth flank clearance b of the ring gear and the planet gears rp Calculating the eccentric value of the planet wheel frame turntable as e p3 ：
Wherein m is r The mass of the inner gear ring of the planet row; omega r Is the angular velocity of the planet row ring gear;
the support error of the planet wheel frame consists of the fit clearance of the inner ring of the support bearing and the input shaft, the fit clearance of the outer ring and the frame and the actual radial clearance of the bearing, the maximum fit clearance of the inner ring and the shaft and the outer ring and the frame are respectively determined based on the basic parameters of the support bearing and the ISO286-1 standard, the actual radial clearance of the bearing is determined based on ISO 5753-1991 and national standard GB/T4604-93, and finally the maximum support error of the planet wheel frame, namely the maximum eccentric amount e caused by the support error of the frame, is determined by superposition calculation p4 。
And the eccentricity of the equivalent turntable of the planet row frame is obtained, so that the subsequent use is convenient.
Further, the method for calculating the total eccentricity of the planet row frame in step S5 is as follows:
total eccentricity e of planet row frame psum The equivalent eccentric amount e of the planet carrier obtained in the step S4 p1 Equivalent eccentricity e of planet row frame p2 The eccentric value of the turntable of the planet wheel frame is e p3 Maximum eccentricity e due to support error of frame p4 Is a vector superposition of (2);
assuming that the eccentric directions caused by the factors are the same, the maximum eccentric amount of the planet row frame is calculated as (e psum ) max The method comprises the following steps:
(e psum ) max ＝e p1 +e p2 +e p3 +e p4
referring to national standard GB/T3748.1-2008, the comprehensive eccentric amount (e psum ) sys The method comprises the following steps:
the operation is simple, and the use is facilitated.
Further, the method for calculating the dynamic unbalance force and dynamic unbalance moment corresponding to each equivalent turntable of the planetary gear mechanism in step S6 is as follows:
calculating dynamic unbalance force f of clutch and planet row frame x 、f y The method comprises the following steps:
wherein m is the equivalent turntable mass; e is the eccentricity of the center of mass and the center of rotation of the turntable; omega is the angular velocity of the turntable;the positive angle between the center of mass of the turntable and the rotation center is X-axis;
calculating dynamic unbalance moment M of clutch and planet row frame x 、M y The method comprises the following steps:
wherein L is the axial width of the turntable; beta is the deflection angle of the inertia shaft and the rotating shaft of the turntable.
And the final dynamic unbalance force parameters are obtained, so that the analysis of the planetary speed change mechanism is facilitated, and the subsequent optimization of the mechanism is performed.
The invention also provides a planet gear unbalanced force assessment method comprising a plurality of non-fixedly connected revolution bodies, which comprises the following steps of:
determining parameters of an optimized planetary transmission mechanism;
optimizing parameters of a planetary speed change mechanism, and determining dynamic unbalance force and dynamic unbalance moment of each equivalent turntable according to the method disclosed by the invention;
and evaluating the influence degree of each optimization measure on the dynamic unbalance force and the dynamic unbalance moment and screening the optimal result.
Therefore, the planetary gear unbalanced force evaluation is realized, the optimized parameters of the planetary speed change mechanism are screened out, the operation is simple, and the use is facilitated.
The invention also provides a planetary gear set dynamic unbalance force determining system comprising a plurality of non-fixed rotators, which comprises a data acquisition module and a processing module, wherein the data acquisition module is used for acquiring basic parameters of a non-fixed rotator component of a planetary gear set speed change mechanism to be calculated, the output end of the data acquisition module is connected with the input end of the processing module, and the processing module executes the method for determining dynamic unbalance force and dynamic unbalance moment corresponding to each equivalent turntable of the planetary gear set speed change mechanism.
By utilizing the system, the unbalanced force of planet gear movement is obtained, and the use is convenient.
The invention also provides a planetary gear unbalanced force evaluation system comprising a plurality of non-fixedly connected revolution bodies, which comprises the planetary gear unbalanced force determination system and an evaluation module, wherein the evaluation module evaluates the influence degree of each optimization measure on dynamic unbalanced force and dynamic unbalanced moment and screens the optimal optimization measure.
By utilizing the evaluation system, optimal optimization measures are screened, and the performance of the planetary speed change mechanism is improved.
Drawings
FIG. 1 is a flow chart of a method for determining a planetary gear imbalance force comprising multiple non-stationary rotors according to the present invention;
FIG. 2 is a simplified illustration of a dual row planetary transmission mechanism input configuration incorporating a method of determining planetary gear imbalance force for multiple non-stationary rotors in accordance with the present invention;
FIG. 3 is a dynamic unbalance force equivalent diagram of the input end of a double-row planetary transmission mechanism incorporating the planetary dynamic unbalance force determination method of multiple non-stationary rotors of the present invention;
FIG. 4 is a schematic diagram of the deflection structure of a CH clutch friction plate incorporating the planetary gear set imbalance force determination method of the present invention with multiple non-stationary rotors;
FIG. 5 is a schematic illustration of the eccentric configuration of a CH clutch friction plate incorporating a planetary gear set imbalance force determination method of multiple non-stationary rotors of the present invention;
FIG. 6 is a schematic diagram of the spline backlash structure of a CH clutch and input shaft of a planetary gear set imbalance force determination method incorporating multiple non-stationary rotors according to the present invention
FIG. 7 is a schematic diagram of a spline tooth configuration of a planetary gear unbalance force determination method incorporating multiple non-stationary rotors of the present invention;
FIG. 8 is a schematic diagram of the positioning error of a planet wheel of a planetary gear set imbalance force determination method incorporating multiple non-stationary rotors of the present invention;
FIG. 9 is a schematic diagram of a planetary gear mass imbalance determination method of the present invention incorporating multiple non-stationary rotors;
fig. 10 is a schematic structural view of a planetary gear set and a planetary gear set tooth side gap in a planetary gear set imbalance force determination method including multiple non-stationary rotors according to the present invention;
fig. 11 is a schematic diagram of the structure of the frame support error of the planetary gear unbalance force determination method including a plurality of non-stationary gyros according to the present invention.
Detailed Description
Embodiments of the present invention are described in detail below, examples of which are illustrated in the accompanying drawings, wherein like or similar reference numerals refer to like or similar elements or elements having like or similar functions throughout. The embodiments described below by referring to the drawings are illustrative only and are not to be construed as limiting the invention.
In the description of the present invention, it should be understood that the terms "longitudinal," "transverse," "upper," "lower," "front," "rear," "left," "right," "vertical," "horizontal," "top," "bottom," "inner," "outer," and the like indicate orientations or positional relationships based on the orientation or positional relationships shown in the drawings, merely to facilitate describing the present invention and simplify the description, and do not indicate or imply that the devices or elements referred to must have a specific orientation, be configured and operated in a specific orientation, and therefore should not be construed as limiting the present invention.
In the description of the present invention, unless otherwise specified and defined, it should be noted that the terms "mounted," "connected," and "coupled" are to be construed broadly, and may be, for example, mechanical or electrical, or may be in communication with each other between two elements, directly or indirectly through intermediaries, as would be understood by those skilled in the art, in view of the specific meaning of the terms described above.
The invention discloses a method for determining unbalanced force of planetary gear set with multiple non-fixed rotators, which comprises the following steps as shown in figure 1:
s1, acquiring basic parameters of a non-fixedly connected revolving body assembly of a planetary gear transmission mechanism to be calculated; as shown in fig. 2 and 3, the non-stationary solid of revolution component includes a clutch, a planet row frame, a planet wheel, an inner gear ring, etc., and basic parameters of the non-stationary solid of revolution component include the number of teeth, a modulus, a reference circle pressure angle of a friction plate and an inner hub in the clutch, the mass of the friction plate and the inner hub, the angular velocity of the friction plate, the mass and the angular velocity of the clutch, the mass and the angular velocity of the planet row frame, the mass and the angular velocity of the inner gear ring, the center distance between the planet wheel and a sun gear, the center distance between the inner gear ring and the planet wheel, the diameter of an inner ring and the outer ring of a support bearing, and tolerance level.
S2, calculating the eccentricity of an equivalent turntable of the clutch oil cylinder caused by the eccentricity of the clutch oil cylinder, the eccentricity of a friction plate and a spline clearance based on basic parameters of the clutch in the non-fixedly connected revolving body assembly;
s3, calculating the total eccentricity of the clutch equivalent turntable according to the eccentricity of the clutch equivalent turntable caused by the eccentricity of the oil cylinder, the eccentricity of the friction plate and the spline clearance, which are determined in the step S2;
s4, calculating the eccentricity of an equivalent turntable of the planet row frame, which is caused by the positioning error, the mass deviation, the tooth side clearance between the inner gear ring and the planet wheel, and the support error of the frame, of the planet row frame in the non-fixedly connected revolving body component based on basic parameters of the planet row frame;
s5, calculating the total eccentricity of the planet row frames according to the eccentricity of each planet row frame caused by the positioning error, the mass deviation, the tooth side gap between the inner gear ring and the planet wheel and the support error of the frames, which are determined in the step S4;
and S6, calculating dynamic unbalance force and dynamic unbalance moment corresponding to each equivalent turntable of the planetary speed change mechanism according to the total eccentricity of the clutch calculated in the step S3 and the total eccentricity of the planetary row frame calculated in the step S5.
The invention estimates the dynamic unbalance force and dynamic unbalance moment of the speed change mechanism, thereby estimating and optimizing the magnitude of the dynamic unbalance force and the dynamic unbalance moment of the relevant parameters of the speed change mechanism, wherein the dynamic unbalance force and the dynamic unbalance moment are an index for measuring the high-speed running performance of the speed change mechanism.
In a preferred embodiment of the present invention, as shown in fig. 4-7, the clutch dynamic imbalance force generation source: deflection of a CH clutch friction plate, eccentricity of the CH clutch friction plate, spline backlash of a CH clutch and an input shaft and spline tooth shape. In step S2, the method for calculating the eccentricity of the equivalent turntable of the clutch oil cylinder caused by the eccentricity of the clutch oil cylinder, the eccentricity of the friction plate and the spline clearance is as follows:
according to basic parameters of clutch cylinder and the eccentric value e of cylinder obtained by experimental measurement 0 The eccentricity equivalent to the clutch turntable is e d1 ：
Wherein m is 0 The mass of the oil cylinder; omega 0 Is the angular velocity of the oil cylinder; m is M d The mass of the clutch turntable; omega d The angular speed of the clutch is the same as the rotational speed of the input shaft;
according to basic parameters of the inner hub and the friction plate, estimating the relation between the eccentric displacement and the spline tooth side gap c by utilizing geometric kinematics, and calculating the initial eccentricity a caused by the tooth side gap on the assumption that the single-side tooth side gap is subjected to Gaussian random distribution:
c i ＝normrnd(μ,σ)
a＝min(a left ,a right )
wherein c i A single-sided tooth flank clearance for the ith tooth; norm represents the generation of random numbers subject to a gaussian distribution (also known as normal distribution); μ is the average value of the tooth flank clearance; σ is the degree of dispersion (standard deviation) of the backlash, the backlash ranging between (μ -3σ, μ+3σ); a, a left The eccentricity of the contact of the left side of the tooth before the right side of the tooth; a, a right The eccentricity of the contact of the right side of the tooth before the left side of the tooth; abs represents the absolute value of the vector in brackets of abs (); n is the number of teeth; alpha is the pressure angle;
by analyzing the eccentricity caused by the backlash under different numbers of teeth, the eccentricity is obtained to be in a decreasing trend along with the increase of the number of teeth, and the larger the number of teeth is, the closer to the backlash is, so when the elastic deformation (such as light load or high rigidity) of the friction plate or spline teeth is not considered, the eccentricity of the friction plate caused by the clearance c is considered to be the same as the clearance, namely e f ＝c；
Based on the derived eccentricity e of the friction plate f Equivalent to the eccentricity e of the clutch disc d2 The method comprises the following steps:
e d2 ＝Nm f e f ω f 2 /M d ω d 2
wherein N is the number of friction plates of the clutch; m is m f Is the mass of the friction plate; omega f Is the angular velocity of the friction plate; omega d The angular speed of the clutch is the same as the rotational speed of the input shaft;
determining the eccentricity e of the spline side gap on the clutch turntable according to the involute spline clearance fit category and the spline tooth tolerance grade in national standard GB/T3748.1-2008 d3 。
More preferably, the method for calculating the total eccentricity of the equivalent rotary disc of the clutch in step S3 is as follows:
total eccentricity e of clutch dsum Calculating the resulting eccentricity e equivalent to the clutch disc for step S2 d1 Eccentric e equivalent to clutch rotor disk d2 And eccentricity e of the spline backlash to the clutch disc d3 If the eccentric direction caused by each factor is the same direction, the maximum eccentric amount of the clutch is calculated as (e dsum ) max The method comprises the following steps:
(e dsum ) max ＝e d1 +e d2 +e d3
calculate the comprehensive eccentric amount (e) dsum ) sys The method comprises the following steps:
in a preferred embodiment of the present invention, as shown in fig. 8-11, the variator planetary imbalance force generation source: planet wheel positioning error, uneven planet wheel quality, inner gear ring and planet wheel tooth side clearance, frame support error. In step S4, the method for calculating the eccentricity of the equivalent turntable of the planet row frame caused by the positioning error, mass deviation, and support error of the ring gear and the planet wheel gear side gap of the frame is as follows:
according to the mass m of the planet wheel 1 And planet wheel position error delta e, calculating equivalent eccentric quantity e of planet row frame p1 The method comprises the following steps:
wherein omega 1 Is the angular velocity of the planet wheel; m is M p Equivalent mass for the planet carrier; omega p Angular velocity for the planet row frame;
calculating the equivalent eccentric quantity e of the planet row frame according to the mass deviation delta m of the planet wheel of the speed change mechanism p2 The method comprises the following steps:
wherein r represents the distance from the center of the planet wheel hole to the rotation center of the planet row frame;
selecting a gear side gap between an inner gear ring and a planet gear according to the center distance of a gear pair, and decentering e caused by the gear side gap between the inner gear ring and the planet gear in a planet gear frame pr About equal to the tooth flank clearance b of the ring gear and the planet gears rp Calculating the eccentric value of the planet wheel frame turntable as e p3 ：
Wherein m is r The mass of the inner gear ring of the planet row; omega r Is the angular velocity of the planet row ring gear;
the support error of the planet wheel frame consists of the fit clearance of the inner ring of the support bearing and the input shaft, the fit clearance of the outer ring and the frame and the actual radial clearance of the bearing, the maximum fit clearance of the inner ring and the shaft and the outer ring and the frame are respectively determined based on the basic parameters of the support bearing and the ISO286-1 standard, the actual radial clearance of the bearing is determined based on ISO 5753-1991 and national standard GB/T4604-93, and finally the maximum support error of the planet wheel frame, namely the maximum eccentric amount e caused by the support error of the frame, is determined by superposition calculation p4 。
In a preferred embodiment of the present invention, the method for calculating the total eccentricity of the planet carrier in step S5 includes:
total eccentricity e of planet row frame psum The equivalent eccentric amount e of the planet carrier obtained in the step S4 p1 Equivalent eccentricity e of planet row frame p2 The eccentric value of the turntable of the planet wheel frame is e p3 Maximum eccentricity e due to support error of frame p4 Is a vector superposition of (2);
assuming that the eccentric directions caused by the factors are the same, the maximum eccentric amount of the planet row frame is calculated as (e psum ) max The method comprises the following steps:
(e psum ) max ＝e p1 +e p2 +e p3 +e p4
referring to national standard GB/T3748.1-2008, the comprehensive eccentric amount (e psum ) sys The method comprises the following steps:
in a preferred scheme of the invention, the method for calculating the dynamic unbalance force and the dynamic unbalance moment corresponding to each equivalent turntable of the planetary speed change mechanism in the step S6 is as follows:
calculating dynamic unbalance force f of clutch and planet row frame x 、f y The method comprises the following steps:
wherein m is the equivalent turntable mass; e is the eccentricity of the center of mass and the center of rotation of the turntable; omega is the angular velocity of the turntable;the positive angle between the center of mass of the turntable and the rotation center is X-axis;
calculating dynamic unbalance moment M of clutch and planet row frame x 、M y The method comprises the following steps:
wherein L is the axial width of the turntable; beta is the deflection angle of the inertia shaft and the rotating shaft of the turntable.
And the final dynamic unbalance force parameters are obtained, so that the analysis of the planetary speed change mechanism is facilitated, and the subsequent optimization of the mechanism is performed.
The invention also provides a planet gear unbalanced force assessment method comprising a plurality of non-fixedly connected revolution bodies, which comprises the following steps of: determining parameters of an optimized planetary transmission mechanism;
optimizing parameters of a planetary speed change mechanism, and determining dynamic unbalance force and dynamic unbalance moment of each equivalent turntable according to the determination method disclosed by the invention;
and evaluating the influence degree of each optimization measure on the dynamic unbalance force and the dynamic unbalance moment and screening the optimal result.
In a preferred scheme of the invention, the method for optimizing parameters of the planetary speed change mechanism to reduce dynamic unbalance force and dynamic unbalance moment corresponding to the equivalent turntable comprises the following steps:
the gap c between the clutch friction plate and the inner hub is reduced, so that the eccentricity e of the equivalent rotary disc of the clutch is reduced d2 . In simulation experiments before and after the gaps between clutch friction plates and inner hubs of a certain double-row planetary speed change mechanism are properly reduced, the equivalent eccentric amount of the improved clutch is obviously reduced, the dynamic unbalance of the improved clutch equivalent turntable is reduced by 7.9% than that before improvement, and the optimization effect is good.
Reducing tolerance level of clutch input shaft and spline teeth to reduce eccentricity e of clutch equivalent turntable d3 . In a simulation experiment after the tolerance level of the key teeth of a certain double-row planetary speed change mechanism is reduced from 6 level to 5 level, the dynamic unbalance force of the clutch after improvement is reduced by 20.21% compared with that before improvement, and the optimization effect is obvious.
Reducing the fit clearance of the inner rings of the support bearings in the planet row so as to reduce the eccentric quantity e caused by the support error of the planet row frame p4 . In a simulation experiment after the fit clearance of the inner rings of the support bearings of a certain double-row planetary speed change mechanism is improved, the dynamic unbalance force of the equivalent turntable of the improved planetary frame is reduced by 3.0% compared with that before improvement, and the optimization effect is slightly good.
Reducing the gap between planetary gear transmission pairs to reduce the eccentric amount e caused by the gap between the inner gear ring and the tooth side of the planetary gear in the planetary gear frame p3 . In simulation experiments before and after the improvement of the clearance of the planetary gear transmission pair of a certain double-row planetary speed change mechanism, the dynamic unbalance force of the equivalent turntable of the improved 1-row and 2-row planetary frame is reduced by 70.57 percent and 79.84 percent compared with that before the improvement, and the optimization effect is obvious.
The invention also provides a planetary gear set dynamic unbalance force determining system comprising a plurality of non-fixed rotators, which comprises a data acquisition module and a processing module, wherein the data acquisition module is used for acquiring basic parameters of a non-fixed rotator component of a planetary gear set speed change mechanism to be calculated, the output end of the data acquisition module is electrically connected with the input end of the processing module, and the processing module executes the determining method to determine dynamic unbalance force and dynamic unbalance moment corresponding to each equivalent turntable of the planetary gear set speed change mechanism. By utilizing the system, the unbalanced force of the planet gear is quickly and accurately obtained, and the system is convenient to use.
The specific data acquisition module can input basic parameters of the bearing through the man-machine interaction module, can acquire all or part of basic parameters of the bearing through a sensor, for example, acquires the number of teeth of a friction plate and an inner hub in a clutch through an image sensor, and supports the diameter of the inner ring and the outer ring of the bearing; acquiring the center distance between the planet wheel and the sun wheel through a distance sensor, the center distance between the inner gear ring and the planet wheel, and the like; and acquiring the angular speed of the friction plate, the angular speed of the clutch, the angular speed of the planet row frame and the angular speed of the inner gear ring by using an angular speed sensor.
The invention also provides a planetary gear unbalanced force evaluation system comprising a plurality of non-fixedly connected revolution bodies, which comprises the planetary gear unbalanced force determination system and the evaluation module, wherein the evaluation module evaluates the influence degree of each optimization measure on dynamic unbalanced force and dynamic unbalanced moment and screens the optimal optimization measure.
In the description of the present specification, a description referring to terms "one embodiment," "some embodiments," "examples," "specific examples," or "some examples," etc., means that a particular feature, structure, material, or characteristic described in connection with the embodiment or example is included in at least one embodiment or example of the present invention. In this specification, schematic representations of the above terms do not necessarily refer to the same embodiments or examples. Furthermore, the particular features, structures, materials, or characteristics described may be combined in any suitable manner in any one or more embodiments or examples.
While embodiments of the present invention have been shown and described, it will be understood by those of ordinary skill in the art that: many changes, modifications, substitutions and variations may be made to the embodiments without departing from the spirit and principles of the invention, the scope of which is defined by the claims and their equivalents.
Claims (10)
1. The method for determining the unbalanced force of the planetary gear set with the multiple non-fixedly connected revolution bodies is characterized by comprising the following steps of:
s1, acquiring basic parameters of a non-fixedly connected revolving body assembly of a planetary gear transmission mechanism to be calculated;
s2, calculating the eccentricity of an equivalent turntable of the clutch oil cylinder caused by the eccentricity of the clutch oil cylinder, the eccentricity of a friction plate and a spline clearance based on basic parameters of the clutch in the non-fixedly connected revolving body assembly;
s3, calculating the total eccentricity of the clutch equivalent turntable according to the eccentricity of the clutch equivalent turntable caused by the eccentricity of the oil cylinder, the eccentricity of the friction plate and the spline clearance, which are determined in the step S2;
s4, calculating the eccentricity of an equivalent turntable of the planet row frame, which is caused by the positioning error, the mass deviation, the tooth side clearance between the inner gear ring and the planet wheel, and the support error of the frame, of the planet row frame in the non-fixedly connected revolving body component based on basic parameters of the planet row frame;
s5, calculating the total eccentricity of the planet row frames according to the eccentricity of each planet row frame caused by the positioning error, the mass deviation, the tooth side gap between the inner gear ring and the planet wheel and the support error of the frames, which are determined in the step S4;
and S6, calculating dynamic unbalance force and dynamic unbalance moment corresponding to each equivalent turntable of the planetary speed change mechanism according to the total eccentricity of the clutch calculated in the step S3 and the total eccentricity of the planetary row frame calculated in the step S5.
2. The method of determining planetary gear unbalance force comprising multiple non-stationary rotors of claim 1, wherein the non-stationary rotor assembly comprises a clutch, a planet gear frame, a planet gear and an annulus gear;
the basic parameters of the non-fixedly connected revolving body component comprise the number of teeth, the modulus and the reference circle pressure angle of a friction plate and an inner hub in a clutch, the mass of the friction plate and the inner hub, the angular velocity of the friction plate, the mass and the angular velocity of the clutch, the mass and the angular velocity of a planet row frame, the mass and the angular velocity of an inner gear ring, the center distance between a planet wheel and a sun wheel, the center distance between the inner gear ring and the planet wheel, the diameter of an inner ring and the outer ring of a support bearing and the tolerance level.
3. The method for determining the unbalanced planetary gear set force comprising a plurality of non-stationary rotors according to claim 1, wherein the method for calculating the eccentricity of the equivalent turntable of the clutch cylinder caused by the eccentricity of the clutch cylinder, the eccentricity of the friction plate and the spline clearance in step S2 is as follows:
according to basic parameters of clutch cylinder and the eccentric value e of cylinder obtained by experimental measurement 0 The eccentricity equivalent to the clutch turntable is e d1 ：
Wherein m is 0 The mass of the oil cylinder; omega 0 Is the angular velocity of the oil cylinder; m is M d The mass of the clutch turntable; omega d The angular speed of the clutch is the same as the rotational speed of the input shaft;
according to basic parameters of the inner hub and the friction plate, estimating the relation between the eccentric displacement and the spline tooth side gap c by utilizing geometric kinematics, and calculating the initial eccentricity a caused by the tooth side gap on the assumption that the single-side tooth side gap is subjected to Gaussian random distribution:
c i ＝normrnd(μ,σ)
a＝min(a left ,a right )
wherein c i A single-sided tooth flank clearance for the ith tooth; norm represents the generation of random numbers subject to gaussian distribution; μ is the average value of the tooth flank clearance; σ is the degree of dispersion of the tooth flank gap, which ranges between (μ -3σ, μ+3σ); a, a left The eccentricity of the contact of the left side of the tooth before the right side of the tooth; a, a right The eccentricity of the contact of the right side of the tooth before the left side of the tooth; abs represents absolute value; n is the number of teeth; alpha is the pressure angle;
by analyzing the eccentricity caused by the backlash under different numbers of teeth, the eccentricity is obtained to be in a decreasing trend along with the increase of the number of teeth, and the larger the number of teeth is, the closer the number of teeth is to the backlash, so when the elastic deformation of the friction plate or the spline teeth is not considered, the eccentricity of the friction plate caused by the clearance c is considered to be the same as the clearance, namely e f ＝c；
Based on the derived eccentricity e of the friction plate f Equivalent to the eccentricity e of the clutch disc d2 The method comprises the following steps:
e d2 ＝Nm f e f ω f 2 /M d ω d 2
wherein N is the number of friction plates of the clutch; m is m f Is the mass of the friction plate; omega f Is the angular velocity of the friction plate; omega d The angular speed of the clutch is the same as the rotational speed of the input shaft;
determining the eccentricity e of the spline side gap on the clutch turntable according to the involute spline clearance fit category and the spline tooth tolerance grade in national standard GB/T3748.1-2008 d3 。
4. The method for determining the unbalanced planetary gear set force comprising a plurality of non-stationary rotors according to claim 1, wherein the method for calculating the total eccentricity of the clutch equivalent turntable in step S3 is as follows:
total eccentricity e of clutch dsum Calculating the resulting eccentricity e equivalent to the clutch disc for step S2 d1 Eccentric e equivalent to clutch rotor disk d2 And eccentricity e of the spline backlash to the clutch disc d3 If the eccentric direction caused by each factor is the same direction, the maximum eccentric amount of the clutch is calculated as (e dsum ) max The method comprises the following steps:
(e dsum ) max ＝e d1 +e d2 +e d3
calculate the comprehensive eccentric amount (e) dsum ) sys The method comprises the following steps:
5. the method for determining the unbalance force of a planetary gear train including a plurality of non-stationary rotors according to claim 1, wherein the method for calculating the eccentricity of the equivalent turntable of the planetary gear train caused by the positioning error of the planetary gear train, the mass deviation, the backlash of the ring gear and the planetary gear, and the supporting error of the frame in step S4 is as follows:
according to the mass m of the planet wheel 1 And planet wheel position error delta e, calculating equivalent eccentric quantity e of planet row frame p1 The method comprises the following steps:
wherein omega 1 Is the angular velocity of the planet wheel; m is M p Equivalent mass for the planet carrier; omega p Angular velocity for the planet row frame;
calculating a planet row frame according to the mass deviation delta m of the planet wheel of the speed change mechanismEquivalent eccentric amount e of rack p2 The method comprises the following steps:
wherein r represents the distance from the center of the planet wheel hole to the rotation center of the planet row frame;
selecting a gear side gap between an inner gear ring and a planet gear according to the center distance of a gear pair, and decentering e caused by the gear side gap between the inner gear ring and the planet gear in a planet gear frame pr About equal to the tooth flank clearance b of the ring gear and the planet gears rp Calculating the eccentric value of the planet wheel frame turntable as e p3 ：
Wherein m is r The mass of the inner gear ring of the planet row; omega r Is the angular velocity of the planet row ring gear;
the support error of the planet wheel frame consists of the fit clearance of the inner ring of the support bearing and the input shaft, the fit clearance of the outer ring and the frame and the actual radial clearance of the bearing, the maximum fit clearance of the inner ring and the shaft and the outer ring and the frame are respectively determined based on the basic parameters of the support bearing and the ISO286-1 standard, the actual radial clearance of the bearing is determined based on ISO 5753-1991 and national standard GB/T4604-93, and finally the maximum support error of the planet wheel frame, namely the maximum eccentric amount e caused by the support error of the frame, is determined by superposition calculation p4 。
6. The method for determining the unbalance force of a planetary gear set including a plurality of non-stationary gyros according to claim 1, wherein the step S5 of calculating the total eccentricity of the planetary gear set frame is:
total eccentricity e of planet row frame psum The equivalent eccentric amount e of the planet carrier obtained in the step S4 p1 Equivalent eccentricity e of planet row frame p2 The eccentric value of the turntable of the planet wheel frame is e p3 Support for frameMaximum eccentricity e caused by error p4 Is a vector superposition of (2);
assuming that the eccentric directions caused by the factors are the same, the maximum eccentric amount of the planet row frame is calculated as (e psum ) max The method comprises the following steps:
(e psum ) max ＝e p1 +e p2 +e p3 +e p4
referring to national standard GB/T3748.1-2008, the comprehensive eccentric amount (e psum ) sys The method comprises the following steps:
7. the method for determining the dynamic unbalance force of a planetary gear set comprising a plurality of non-stationary rotors according to claim 1, wherein the step S6 is a method for calculating the dynamic unbalance force and the dynamic unbalance moment corresponding to each equivalent turntable of the planetary gear set as follows:
calculating dynamic unbalance force f of clutch and planet row frame x 、f y The method comprises the following steps:
wherein m is the equivalent turntable mass; e is the eccentricity of the center of mass and the center of rotation of the turntable; omega is the angular velocity of the turntable;the positive angle between the center of mass of the turntable and the rotation center is X-axis;
calculating dynamic unbalance moment M of clutch and planet row frame x 、M y The method comprises the following steps:
wherein L is the axial width of the turntable; beta is the deflection angle of the inertia shaft and the rotating shaft of the turntable.
8. The planetary gear unbalance force evaluation method comprising a plurality of non-fixedly connected revolution bodies is characterized by comprising the following steps of:
determining parameters of an optimized planetary transmission mechanism;
optimizing parameters of a planetary speed change mechanism and determining dynamic unbalance force and dynamic unbalance moment of each equivalent turntable according to the method of one of claims 1 to 7;
and evaluating the influence degree of each optimization measure on the dynamic unbalance force and the dynamic unbalance moment and screening the optimal result.
9. A planetary gear unbalance force determining system comprising a plurality of non-stationary rotators, which is characterized by comprising a data acquisition module and a processing module, wherein the data acquisition module is used for acquiring basic parameters of a non-stationary rotators assembly of a planetary gear shift mechanism to be calculated, the output end of the data acquisition module is connected with the input end of the processing module, and the processing module executes the method of one of claims 1-7 to determine dynamic unbalance force and dynamic unbalance moment corresponding to each equivalent turntable of the planetary gear shift mechanism.
10. A planetary gear unbalance force evaluation system comprising multiple non-stationary rotors, comprising the planetary gear unbalance force determination system of claim 9, and an evaluation module that evaluates the extent of influence of each optimization measure on dynamic unbalance force and dynamic unbalance moment and screens out the optimal optimization measure.
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| CN117644471A (en) * | 2024-01-29 | 2024-03-05 | 北京特思迪半导体设备有限公司 | Method for precisely adjusting eccentricity of eccentric driving mechanism |
| CN117644471B (en) * | 2024-01-29 | 2024-04-16 | 北京特思迪半导体设备有限公司 | Method for precisely adjusting eccentricity of eccentric driving mechanism |
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