CN113104017A - Hydraulic power-assisted mode pump displacement control method based on optimal driving force distribution - Google Patents

Hydraulic power-assisted mode pump displacement control method based on optimal driving force distribution Download PDF

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CN113104017A
CN113104017A CN202110532602.1A CN202110532602A CN113104017A CN 113104017 A CN113104017 A CN 113104017A CN 202110532602 A CN202110532602 A CN 202110532602A CN 113104017 A CN113104017 A CN 113104017A
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hydraulic
wheel
driving force
speed
pump
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曾小华
吴梓乔
宋大凤
杨东坡
张轩铭
高福旺
钱琦峰
王诗元
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Jilin University
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    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
    • B60W10/00—Conjoint control of vehicle sub-units of different type or different function
    • B60W10/04—Conjoint control of vehicle sub-units of different type or different function including control of propulsion units
    • B60W10/06—Conjoint control of vehicle sub-units of different type or different function including control of propulsion units including control of combustion engines
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
    • B60W10/00—Conjoint control of vehicle sub-units of different type or different function
    • B60W10/04—Conjoint control of vehicle sub-units of different type or different function including control of propulsion units
    • B60W10/08—Conjoint control of vehicle sub-units of different type or different function including control of propulsion units including control of electric propulsion units, e.g. motors or generators
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
    • B60W20/00—Control systems specially adapted for hybrid vehicles
    • B60W20/10—Controlling the power contribution of each of the prime movers to meet required power demand
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
    • B60W40/00—Estimation or calculation of non-directly measurable driving parameters for road vehicle drive control systems not related to the control of a particular sub unit, e.g. by using mathematical models
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
    • B60W2710/00—Output or target parameters relating to a particular sub-units
    • B60W2710/06—Combustion engines, Gas turbines
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60W—CONJOINT CONTROL OF VEHICLE SUB-UNITS OF DIFFERENT TYPE OR DIFFERENT FUNCTION; CONTROL SYSTEMS SPECIALLY ADAPTED FOR HYBRID VEHICLES; ROAD VEHICLE DRIVE CONTROL SYSTEMS FOR PURPOSES NOT RELATED TO THE CONTROL OF A PARTICULAR SUB-UNIT
    • B60W2710/00—Output or target parameters relating to a particular sub-units
    • B60W2710/08—Electric propulsion units
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
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Abstract

The invention discloses a hydraulic power-assisted mode pump displacement control method based on optimal driving force distribution.

Description

Hydraulic power-assisted mode pump displacement control method based on optimal driving force distribution
Technical Field
The invention relates to a displacement control method of a hydraulic booster mode pump based on optimal driving force distribution, in particular to a displacement control method of a variable displacement pump of a hub hydraulic system.
Background
Different from an oil-electricity hybrid power system, the hub hydraulic hybrid power system is a typical strong-nonlinearity parameter time-varying electromechanical-hydraulic coupling control system, the response characteristic difference of a front wheel hydraulic transmission part and a middle-rear wheel mechanical transmission part is obvious, and the hub hydraulic hybrid power system is influenced by the characteristics of complex operation conditions of heavy commercial vehicles and large-range load variation, the dynamic control quality of the hub hydraulic hybrid power system is difficult to guarantee, on one hand, the driving force control between the hydraulic transmission system and the mechanical transmission system is easy to generate interference, and the exertion of the power-assisting function of the system is influenced; on the other hand, the inherent nonlinearity problem of the hydraulic system is also easy to cause the control of the hydraulic actuator to lag or overshoot, which results in the slow response of the hydraulic system or the generation of large pressure impact, and influences the dynamic control performance of the system.
At present, the hub hydraulic hybrid power system is still in a starting stage in the domestic research, the key technical theory researches such as careful scheme optimization and core control algorithm development around the system still have higher theoretical significance and application value, and the following technical difficulty problems still exist in the hub hydraulic hybrid power system at present: the wheel hub hydraulic hybrid power system solves the problem of the coordinated distribution of the engine power between a front wheel hydraulic system and a middle and rear wheel mechanical system in a closed hydraulic loop pump power-assisted mode, mainly shows in the control process of a hydraulic variable pump, on one hand, the wheel speed of a front wheel and the wheel speed of a middle and rear wheel must be coordinated, interference cannot be generated, otherwise the system power-assisted effect is greatly weakened, and even the system power-assisted effect can play a role of blocking; on the other hand, on the premise of ensuring the assistance of the hydraulic system, the traction efficiency of the whole vehicle transmission system is improved as much as possible.
The invention provides a hub hydraulic drive variable pump control method, which can ensure that the sliding efficiency of a whole vehicle is optimal when the vehicle passes through a low-attachment road surface or a large-gradient road surface. The invention discloses a method for controlling the displacement of a main pump in a main winch confluence lifting process of a rotary drilling rig, wherein the Chinese patent publication number is CN105502191A, the publication number is 2016, 4, 20, and the invention provides that the dropping rate is used as a judgment parameter of the dropping speed state, when the dropping rate is smaller than a set value, no adjustment is carried out, and when the dropping rate is larger than the set value, the displacement of the main pump is controlled by adjusting the input current value of a proportional pressure reducing valve through a real-time PID.
In summary, the existing patents on the aspect of controlling the displacement of the hub hydraulic variable pump only perform control by a simple method of independently considering slip or speed following, and cannot analyze the characteristics of the internal hydraulic and mechanical systems, so that the result is over-ideal, and the actual control effect is limited. Therefore, it is necessary to provide a perfect and reliable displacement control method for a hub hydraulic variable pump to make up for the deficiencies of the prior art, comprehensively consider the characteristics of the mechanical system and the hydraulic system, ensure the coordinated distribution of the engine power between the front wheel hydraulic system and the middle and rear wheel mechanical systems in the power-assisted mode of the closed hydraulic circuit pump, enable the hydraulic execution component to respond well to the optimized control target in each mode, and improve the dynamic control quality of the hub hydraulic hybrid power system.
Disclosure of Invention
The invention aims to solve the problem of coordinated distribution of engine power between a front wheel hydraulic system and a middle and rear wheel mechanical system in a closed hydraulic loop pump power-assisted mode of a hub hydraulic hybrid power system, and provides a pump displacement control optimization method in the closed loop pump power-assisted mode of the hub hydraulic hybrid power system.
In order to solve the technical problems, the invention is realized by adopting the following technical scheme:
the method comprises the following steps that firstly, required torque and power of a front wheel hydraulic hub motor are obtained through solving based on a hydraulic closed loop flow consistency principle according to the configuration of a whole vehicle and basic parameters of components, and an optimal displacement control target of a variable displacement pump is further obtained;
according to the principle of flow consistency of a hydraulic closed loop, the flow of two hub hydraulic motors in a hydraulic power-assisted mode is equal to the output flow of a variable pump:
ωpβVpmaxηpvηvv＝2ωmVm/ηmv (1)
in the formula, ωp、ωmRespectively representing the rotational speed of the hydraulic variable pump and the rotational speed, eta, of the hub hydraulic motorvvThe efficiency loss of the meter type hydraulic control valve group and the pipeline, beta represents the opening degree of a swash plate of the hydraulic variable pump, and can also be understood as beta represents the displacement of the hydraulic variable pump, and VpmaxIs the maximum displacement, V, of the hydraulic variable displacement pumpmIs the displacement of the hub hydraulic motor, etapvIndicating the volumetric efficiency, η, of a hydraulic variable displacement pumpmvIs the volumetric efficiency of the hydraulic motor;
at this time, the rotation speed omega of the hub hydraulic motormWith the speed omega of the hydraulic variable pumppSatisfies the relationship:
Figure BDA0003068485480000021
simultaneously, the sum of the output torques of the two front wheel hub hydraulic motors is calculated:
Figure BDA0003068485480000022
in the formula, Δ P represents the pressure difference between the oil inlet and the oil outlet of the hydraulic motor passing through the hub, ηmmIndicating the mechanical efficiency of the hub hydraulic motor;
further, the output power P of the hydraulic motor of the wheel hub at the wheel is obtainedm：
Figure BDA0003068485480000023
The hub hydraulic hybrid power system is characterized in that an engine and a variable pump are connected through a PTO (power take off), so that the following relation is satisfied between the engine speed and the variable pump speed:
ωp＝ωe/ip (5)
in the formula, ωeIndicating engine speed, ipRepresenting a PTO speed ratio;
according to the hydraulic system flow consistency requirement, the relation between the hub hydraulic motor rotating speed and the front wheel speed and the relation between the engine rotating speed and the middle and rear wheel speed shown in the formula (1) to the formula (5), the variable pump target displacement meeting the hydraulic system flow consistency in the boosting mode is further deduced:
Figure BDA0003068485480000031
in the formula igIndicating the current gear ratio, i, of the transmission0Representing main-reducer speed ratio, ωfAnd omegarRespectively representing the wheel speeds of the front wheel and the middle and rear wheels of the vehicle;
the target displacement of the variable displacement pump in the power-assisted mode of the closed hydraulic circuit pump mainly depends on the wheel speed and the speed ratio of the transmission, and is also influenced by the volumetric efficiency of a hydraulic system; under a certain working condition, the wheel speed of a rear wheel in the vehicle is mainly determined by the rotating speed of an engine; under the condition that the working point of the engine is fixed, the displacement of the variable displacement pump is controlled to be increased, and the wheel speed of the front wheel can be controlled to be increased; therefore, if the optimal front wheel rotating speed in the mode can be determined, the optimal displacement control target of the variable displacement pump can be determined;
converting the optimal target displacement problem of the variable pump into an optimal control problem of the rotating speed of a front wheel of the vehicle, defining the traction efficiency of the whole vehicle based on the slip loss of the vehicle and the distribution coefficient of the driving force of the whole vehicle, and determining factors influencing the traction efficiency of the whole vehicle;
according to the analysis in the first step, if the optimal front wheel rotating speed in the mode can be determined, the optimal displacement control target of the variable pump can be determined, and the optimal displacement control problem of the variable pump is converted into the control problem of the optimal rotating speed of the front wheel of the vehicle;
firstly, the traction efficiency of the system is defined according to the slip loss generated by the wheel slip during the running process of the all-wheel drive vehicle:
Figure BDA0003068485480000032
in the formula etasRepresenting the system traction efficiency, sfAnd srRespectively representing the slip rates, v, of the front and middle rear wheels of the vehiclefAnd vrRepresenting the speed of the front and middle rear wheels of the vehicle, respectively, Ff、FmAnd FrRespectively representing the driving forces of the front, middle and rear axles of the vehicle;
further, a whole vehicle driving force distribution coefficient K is definedd：
Kd＝(Fm+Fr)/(Ff+Fm+Fr) (8)
The practical meaning of the distribution coefficient is the proportion of the driving force of the mechanical transmission path to the total driving force; then, combining the relationship of the rotation speeds of the wheels of the vehicle, the following is obtained:
v＝vf(1-sf)＝vm(1-sm)＝vr(1-sr) (9)
the traction efficiency of a vehicle in all-wheel drive can be further expressed as:
Figure BDA0003068485480000033
according to the formula (10), in the closed hydraulic circuit boosting mode, the system traction efficiency mainly depends on the slip ratio of each wheel and the distribution condition of the whole vehicle driving force in the front wheel hydraulic path and the middle and rear wheel mechanical path of the vehicle;
determining the optimal traction efficiency based on the driving force distribution coefficient; according to the formula of the traction efficiency of the whole vehicle determined in the step two, the wheel slip rate and the driving of the whole vehicle are basedThe force distribution coefficient is further subjected to derivation on the formula (10) to obtain the system traction efficiency etasTo driving force distribution coefficient KdFirst and second partial derivatives of (c):
Figure BDA0003068485480000041
when the first partial derivative f is expressed by equation (11)1When the driving force distribution coefficient is equal to 0, the corresponding driving force distribution coefficient can enable the traction efficiency of the system to obtain an extreme value, and the slip ratios of the front wheels and the middle and rear wheels have three states: s f1 or sr1 or sf＝sr(ii) a Considering that the wheel slip rate cannot reach 100% in the normal running process of the vehicle, s is equal to and only equal tof＝srIn time, the first-order partial derivative is zero, which is a necessary condition for obtaining the optimal traction efficiency of the whole vehicle; but due to its second partial derivative f2During the change of the slip ratio of the front wheel and the middle and rear wheels, f can not be ensured2If < 0 is always true, s cannot be guaranteedf＝srThe system traction efficiency is maximum, so the extreme point of the formula (11) is further analyzed by combining the change of the driving force distribution coefficient and the states of the speeds of the front wheel and the middle and rear wheels;
firstly, under the static condition that the slip rates of front and rear wheels are not changed; current wheel speed<When the rear wheel is rotating, sf＜srI.e. first partial derivatives f1< 0 is always true; at this time, the distribution coefficient K is varieddThe driving force of a mechanical path is reduced, the driving force of a hydraulic path is increased, and the traction efficiency of the system is gradually increased; and the current wheel speed>When the rear wheel is rotating, sf＞srI.e. first partial derivatives f1If more than 0 is always true; at this time, the distribution coefficient K is varieddThe driving force of a mechanical path is reduced, the driving force of a hydraulic path is increased, and the traction efficiency of the system is gradually reduced; it can be seen that in this static case, when K isd1 or KdWhen the traction efficiency is equal to 0, the system traction efficiency reaches a maximum value;
in the actual running process of the vehicle, the driving force distribution coefficient and the slip ratio of the wheels are not isolated, different driving force distribution ratios have the influence on the change of the slip ratio of the wheels, and the slip ratio of the wheels is increased along with the increase of the driving force, so that the vehicle can not obtain the optimal traction efficiency under the condition that the driving force distribution coefficient is extreme;
therefore, if the front wheel speed is changed during the change of the driving force distribution ratio<Rear wheel speed, with distribution coefficient KdThe driving force of the mechanical path is reduced, the driving force of the hydraulic path is increased, the slip rate of the front wheel is gradually increased, the slip rate of the middle wheel and the rear wheel is also gradually reduced, the traction efficiency of the system is gradually increased, but the corresponding K is obtaineddThe traction efficiency of the system is gradually reduced under the condition of 0; at distribution coefficient KdDuring the reduction process, the wheel speed of the front wheel may exceed that of the rear wheel, and at the moment, the traction efficiency of the system follows KdDecrease to further decrease; it can be seen that with the change of the front and rear wheel speeds, there is always a driving force distribution point to make the traction efficiency of the system reach the maximum value in the dynamic change process, and the point is that the corresponding slip ratio satisfies sf＝srAt the point where the optimum tractive efficiency achievable by the system is ηs,max＝1-sf＝1-sr(ii) a Furthermore, neglecting the difference between the front wheel and the middle and rear wheel, the wheel speed of the front wheel is equal to the wheel speed of the middle and rear wheel;
in summary, when the system works in the closed hydraulic circuit pump boosting mode, the distribution condition of the driving force in the front wheel hydraulic path and the middle and rear wheel mechanical paths should be adjusted as much as possible, so that the wheel speed of the front wheel and the wheel speed of the middle and rear wheels can be followed, and when the control target is reached, the optimal pump displacement target in the mode is as follows:
Figure BDA0003068485480000042
hydraulic variable displacement pump volumetric efficiency etapvAnd the volumetric efficiency eta of the hub hydraulic motormvThe calculation formulas of (A) and (B) are respectively as follows:
Figure BDA0003068485480000051
Figure BDA0003068485480000052
wherein μ represents the dynamic viscosity of the hydraulic oil, CpsRepresents the laminar leakage coefficient, C, of the hydraulic pumpmsRepresenting a laminar leakage coefficient of the hydraulic motor;
and (5) obtaining a final expression form of the target displacement control law of the variable displacement pump according to the expressions (12) to (14):
Figure BDA0003068485480000053
according to equation (15), the target control displacement of the variable displacement pump in the closed hydraulic circuit pump assist mode is composed of two parts: one is the fixed value part associated with the gear,
Figure BDA0003068485480000054
the second is an efficiency compensation part caused by relevant factors such as rotating speed, working pressure and the like,
Figure BDA0003068485480000055
when the vehicle runs at a certain speed, the gear is fixed, and the fixed displacement part beta is usedgOnly depends on the gear of the speed changer and can be obtained by calculation according to the parameters of the whole vehicle, and the efficiency compensation part betaηAnd finally calculating to obtain a pump displacement control target in the closed hydraulic circuit pump boosting mode according to different engine rotating speeds and system pressures related to the system pressure and the engine rotating speed.
Compared with the prior art, the invention has the beneficial effects that: the invention fully considers the mechanical and hydraulic system characteristics of the hub hydraulic hybrid power system, carries out pump displacement control based on the optimal driving force of the system and can realize the comprehensive optimal control effect; in addition, the control method is simple in control flow and easy to implement in engineering.
Drawings
FIG. 1 is a structural diagram of a hub hydraulic hybrid power system according to the present invention;
FIG. 2 is an overall flow chart of the displacement control of the variable displacement pump of the hub hydraulic hybrid power system according to the invention;
fig. 3 shows the calculation solution of the displacement of the hydraulic power-assisted pump based on the optimal driving force distribution according to the invention.
Detailed Description
The present invention will be described in detail below with reference to the accompanying drawings.
Referring to the attached figure 1, the invention is applied to a front axle hub hydraulic system driven and a hub hydraulic hybrid power system driven by a middle-rear axle engine;
referring to the attached figure 2, the invention provides a hydraulic power-assisted mode pump displacement control method based on optimal driving force distribution, which is based on a hub hydraulic hybrid power system, carries out hydraulic-mechanical path characteristic analysis of a finished vehicle on the basis of existing finished vehicle components and basic parameters, considers finished vehicle driving force distribution, and carries out variable pump displacement optimal control.
The method specifically comprises the following steps:
the method comprises the following steps that firstly, required torque and power of a front wheel hydraulic hub motor are obtained through solving based on a hydraulic closed loop flow consistency principle according to the configuration of a whole vehicle and basic parameters of components, and an optimal displacement control target of a variable displacement pump is further obtained;
according to the principle of flow consistency of a hydraulic closed loop, the flow of two hub hydraulic motors in a hydraulic power-assisted mode is equal to the output flow of a variable pump:
ωpβVpmaxηpvηvv＝2ωmVm/ηmv (1)
in the formula, ωp、ωmRespectively representing the rotational speed of the hydraulic variable pump and the rotational speed, eta, of the hub hydraulic motorvvThe efficiency loss of the meter type hydraulic control valve group and the pipeline, beta represents the opening degree of a swash plate of the hydraulic variable pump, and can also be understood as beta represents the displacement of the hydraulic variable pump, and VpmaxIs the maximum displacement, V, of the hydraulic variable displacement pumpmFor wheel hub hydraulicsDisplacement of the motor, ηpvIndicating the volumetric efficiency, η, of a hydraulic variable displacement pumpmvIs the volumetric efficiency of the hydraulic motor;
at this time, the rotation speed omega of the hub hydraulic motormWith the speed omega of the hydraulic variable pumppSatisfies the relationship:
Figure BDA0003068485480000061
simultaneously, the sum of the output torques of the two front wheel hub hydraulic motors is calculated:
Figure BDA0003068485480000062
in the formula, Δ P represents the pressure difference between the oil inlet and the oil outlet of the hydraulic motor passing through the hub, ηmmIndicating the mechanical efficiency of the hub hydraulic motor; further, the output power P of the hydraulic motor of the wheel hub at the wheel is obtainedm：
Figure BDA0003068485480000063
The hub hydraulic hybrid power system is characterized in that an engine and a variable pump are connected through a PTO (power take off), so that the following relation is satisfied between the engine speed and the variable pump speed:
ωp＝ωe/ip (5)
in the formula, ωeIndicating engine speed, ipRepresenting a PTO speed ratio;
according to the hydraulic system flow consistency requirement, the relation between the hub hydraulic motor rotating speed and the front wheel speed and the relation between the engine rotating speed and the middle and rear wheel speed shown in the formula (1) to the formula (5), the variable pump target displacement meeting the hydraulic system flow consistency in the boosting mode is further deduced:
Figure BDA0003068485480000064
in the formula igIndicating the current gear ratio, i, of the transmission0Representing main-reducer speed ratio, ωfAnd omegarRespectively representing the wheel speeds of the front wheel and the middle and rear wheels of the vehicle;
the target displacement of the variable displacement pump in the power-assisted mode of the closed hydraulic circuit pump mainly depends on the wheel speed and the speed ratio of the transmission, and is also influenced by the volumetric efficiency of a hydraulic system; under a certain working condition, the wheel speed of a rear wheel in the vehicle is mainly determined by the rotating speed of an engine; under the condition that the working point of the engine is fixed, the displacement of the variable displacement pump is controlled to be increased, and the wheel speed of the front wheel can be controlled to be increased; therefore, if the optimal front wheel rotating speed in the mode can be determined, the optimal displacement control target of the variable displacement pump can be determined;
converting the optimal target displacement problem of the variable pump into an optimal control problem of the rotating speed of a front wheel of the vehicle, defining the traction efficiency of the whole vehicle based on the slip loss of the vehicle and the distribution coefficient of the driving force of the whole vehicle, and determining factors influencing the traction efficiency of the whole vehicle;
according to the analysis in the first step, if the optimal front wheel rotating speed in the mode can be determined, the optimal displacement control target of the variable pump can be determined, and the optimal displacement control problem of the variable pump is converted into the control problem of the optimal rotating speed of the front wheel of the vehicle;
firstly, the traction efficiency of the system is defined according to the slip loss generated by the wheel slip during the running process of the all-wheel drive vehicle:
Figure BDA0003068485480000071
in the formula etasRepresenting the system traction efficiency, sfAnd srRespectively representing the slip rates, v, of the front and middle rear wheels of the vehiclefAnd vrRepresenting the speed of the front and middle rear wheels of the vehicle, respectively, Ff、FmAnd FrRespectively representing the driving forces of the front, middle and rear axles of the vehicle;
further, a whole vehicle driving force distribution coefficient K is definedd：
Kd＝(Fm+Fr)/(Ff+Fm+Fr) (8)
The practical meaning of the distribution coefficient is the proportion of the driving force of the mechanical transmission path to the total driving force; then, combining the relationship of the rotation speeds of the wheels of the vehicle, the following is obtained:
v＝vf(1-sf)＝vm(1-sm)＝vr(1-sr) (9)
the traction efficiency of a vehicle in all-wheel drive can be further expressed as:
Figure BDA0003068485480000072
according to the formula (10), in the closed hydraulic circuit boosting mode, the system traction efficiency mainly depends on the slip ratio of each wheel and the distribution condition of the whole vehicle driving force in the front wheel hydraulic path and the middle and rear wheel mechanical path of the vehicle;
determining the optimal traction efficiency based on the driving force distribution coefficient; according to the whole vehicle traction efficiency formula determined in the step two, the formula (10) is further subjected to derivation based on the wheel slip rate and the whole vehicle driving force distribution coefficient to obtain the system traction efficiency etasTo driving force distribution coefficient KdFirst and second partial derivatives of (c):
Figure BDA0003068485480000073
when the first partial derivative f is expressed by equation (11)1When the driving force distribution coefficient is equal to 0, the corresponding driving force distribution coefficient can enable the traction efficiency of the system to obtain an extreme value, and the slip ratios of the front wheels and the middle and rear wheels have three states: sf1 or s r1 or sf＝sr(ii) a Considering that the wheel slip rate cannot reach 100% in the normal running process of the vehicle, s is equal to and only equal tof＝srIn time, the first-order partial derivative is zero, which is a necessary condition for obtaining the optimal traction efficiency of the whole vehicle; but due to its second partial derivative f2At the slip rate of the front and middle rear wheelsDuring the change, f cannot be guaranteed2If < 0 is always true, s cannot be guaranteedf＝srThe system traction efficiency is maximum, so the extreme point of the formula (11) is further analyzed by combining the change of the driving force distribution coefficient and the states of the speeds of the front wheel and the middle and rear wheels;
firstly, under the static condition that the slip rates of front and rear wheels are not changed; current wheel speed<When the rear wheel is rotating, sf＜srI.e. first partial derivatives f1< 0 is always true; at this time, the distribution coefficient K is varieddThe driving force of a mechanical path is reduced, the driving force of a hydraulic path is increased, and the traction efficiency of the system is gradually increased; and the current wheel speed>When the rear wheel is rotating, sf＞srI.e. first partial derivatives f1If more than 0 is always true; at this time, the distribution coefficient K is varieddThe driving force of a mechanical path is reduced, the driving force of a hydraulic path is increased, and the traction efficiency of the system is gradually reduced; it can be seen that in this static case, when K isd1 or KdWhen the traction efficiency is equal to 0, the system traction efficiency reaches a maximum value;
in the actual running process of the vehicle, the driving force distribution coefficient and the slip ratio of the wheels are not isolated, different driving force distribution ratios have the influence on the change of the slip ratio of the wheels, and the slip ratio of the wheels is increased along with the increase of the driving force, so that the vehicle can not obtain the optimal traction efficiency under the condition that the driving force distribution coefficient is extreme;
therefore, if the front wheel speed is changed during the change of the driving force distribution ratio<Rear wheel speed, with distribution coefficient KdThe driving force of the mechanical path is reduced, the driving force of the hydraulic path is increased, the slip rate of the front wheel is gradually increased, the slip rate of the middle wheel and the rear wheel is also gradually reduced, the traction efficiency of the system is gradually increased, but the corresponding K is obtaineddThe traction efficiency of the system is gradually reduced under the condition of 0; at distribution coefficient KdDuring the reduction process, the wheel speed of the front wheel may exceed that of the rear wheel, and at the moment, the traction efficiency of the system follows KdDecrease to further decrease; it can be seen that as the wheel speed changes, there is always a driving force distribution point to make the system tractionThe efficiency reaches the maximum value in the dynamic change process, and the point is that the corresponding slip ratio satisfies sf＝srAt the point where the optimum tractive efficiency achievable by the system is ηs,max＝1-sf＝1-sr(ii) a Furthermore, neglecting the difference between the front wheel and the middle and rear wheel, the wheel speed of the front wheel is equal to the wheel speed of the middle and rear wheel;
in summary, when the system works in the closed hydraulic circuit pump boosting mode, the distribution condition of the driving force in the front wheel hydraulic path and the middle and rear wheel mechanical paths should be adjusted as much as possible, so that the wheel speed of the front wheel and the wheel speed of the middle and rear wheels can be followed, and when the control target is reached, the optimal pump displacement target in the mode is as follows:
Figure BDA0003068485480000081
hydraulic variable displacement pump volumetric efficiency etapvAnd the volumetric efficiency eta of the hub hydraulic motormvThe calculation formulas of (A) and (B) are respectively as follows:
Figure BDA0003068485480000082
Figure BDA0003068485480000083
wherein μ represents the dynamic viscosity of the hydraulic oil, CpsRepresents the laminar leakage coefficient, C, of the hydraulic pumpmsRepresenting a laminar leakage coefficient of the hydraulic motor;
and (5) obtaining a final expression form of the target displacement control law of the variable displacement pump according to the expressions (12) to (14):
Figure BDA0003068485480000091
according to equation (15), the target control displacement of the variable displacement pump in the closed hydraulic circuit pump assist mode is composed of two parts: one is the fixed value part associated with the gear,
Figure BDA0003068485480000092
the second is an efficiency compensation part caused by relevant factors such as rotating speed, working pressure and the like,
Figure BDA0003068485480000093
when the vehicle runs at a certain speed, the gear is fixed, and the fixed displacement part beta is usedgOnly depends on the gear of the speed changer and can be obtained by calculation according to the parameters of the whole vehicle, and the efficiency compensation part betaηAnd finally calculating to obtain a pump displacement control target in the closed hydraulic circuit pump boosting mode according to different engine rotating speeds and system pressures related to the system pressure and the engine rotating speed.
Beta corresponding to different gears of certain hub hydraulic hybrid power systemgComprises the following steps:
Figure BDA0003068485480000094
referring to the attached figure 3, after the system pressure, the engine speed and the current gear are comprehensively determined, the pump displacement control target in the closed hydraulic circuit pump power assisting mode of the hub hydraulic hybrid power system can be finally calculated.
Parts which are not described in the invention can be realized by adopting or referring to the prior art.
The above description is only an example of the present invention and should not be taken as limiting the invention, and any modifications, equivalents, improvements and the like made within the spirit and principle of the present invention should be included in the scope of the present invention.

Claims (1)

1. A method of controlling displacement of a hydraulic assist mode pump based on optimal power allocation, comprising the steps of:
the method comprises the following steps that firstly, required torque and power of a front wheel hydraulic hub motor are obtained through solving based on a hydraulic closed loop flow consistency principle according to the configuration of a whole vehicle and basic parameters of components, and an optimal displacement control target of a variable displacement pump is further obtained;
according to the principle of flow consistency of a hydraulic closed loop, the flow of two hub hydraulic motors in a hydraulic power-assisted mode is equal to the output flow of a variable pump:
ωpβVpmaxηpvηvv＝2ωmVm/ηmv (1)
in the formula, ωp、ωmRespectively representing the rotational speed of the hydraulic variable pump and the rotational speed, eta, of the hub hydraulic motorvvThe efficiency loss of the meter type hydraulic control valve group and the pipeline, beta represents the opening degree of a swash plate of the hydraulic variable pump, and can also be understood as beta represents the displacement of the hydraulic variable pump, and VpmaxIs the maximum displacement, V, of the hydraulic variable displacement pumpmIs the displacement of the hub hydraulic motor, etapvIndicating the volumetric efficiency, η, of a hydraulic variable displacement pumpmvIs the volumetric efficiency of the hydraulic motor;
at this time, the rotation speed omega of the hub hydraulic motormWith the speed omega of the hydraulic variable pumppSatisfies the relationship:
Figure FDA0003068485470000011
simultaneously, the sum of the output torques of the two front wheel hub hydraulic motors is calculated:
Figure FDA0003068485470000012
in the formula, Δ P represents the pressure difference between the oil inlet and the oil outlet of the hydraulic motor passing through the hub, ηmmIndicating the mechanical efficiency of the hub hydraulic motor;
further, the output power P of the hydraulic motor of the wheel hub at the wheel is obtainedm：
Figure FDA0003068485470000013
The hub hydraulic hybrid power system is characterized in that an engine and a variable pump are connected through a PTO (power take off), so that the following relation is satisfied between the engine speed and the variable pump speed:
ωp＝ωe/ip (5)
in the formula, ωeIndicating engine speed, ipRepresenting a PTO speed ratio;
according to the hydraulic system flow consistency requirement, the relation between the hub hydraulic motor rotating speed and the front wheel speed and the relation between the engine rotating speed and the middle and rear wheel speed shown in the formula (1) to the formula (5), the variable pump target displacement meeting the hydraulic system flow consistency in the boosting mode is further deduced:
Figure FDA0003068485470000014
in the formula igIndicating the current gear ratio, i, of the transmission0Representing main-reducer speed ratio, ωfAnd omegarRespectively representing the wheel speeds of the front wheel and the middle and rear wheels of the vehicle;
the target displacement of the variable displacement pump in the power-assisted mode of the closed hydraulic circuit pump mainly depends on the wheel speed and the speed ratio of the transmission, and is also influenced by the volumetric efficiency of a hydraulic system; under a certain working condition, the wheel speed of a rear wheel in the vehicle is mainly determined by the rotating speed of an engine; under the condition that the working point of the engine is fixed, the displacement of the variable displacement pump is controlled to be increased, and the wheel speed of the front wheel can be controlled to be increased; therefore, if the optimal front wheel rotating speed in the mode can be determined, the optimal displacement control target of the variable displacement pump can be determined;
converting the optimal target displacement problem of the variable pump into an optimal control problem of the rotating speed of a front wheel of the vehicle, defining the traction efficiency of the whole vehicle based on the slip loss of the vehicle and the distribution coefficient of the driving force of the whole vehicle, and determining factors influencing the traction efficiency of the whole vehicle;
according to the analysis in the first step, if the optimal front wheel rotating speed in the mode can be determined, the optimal displacement control target of the variable pump can be determined, and the optimal displacement control problem of the variable pump is converted into the control problem of the optimal rotating speed of the front wheel of the vehicle;
firstly, the traction efficiency of the system is defined according to the slip loss generated by the wheel slip during the running process of the all-wheel drive vehicle:
Figure FDA0003068485470000021
in the formula etasRepresenting the system traction efficiency, sfAnd srRespectively representing the slip rates, v, of the front and middle rear wheels of the vehiclefAnd vrRepresenting the speed of the front and middle rear wheels of the vehicle, respectively, Ff、FmAnd FrRespectively representing the driving forces of the front, middle and rear axles of the vehicle;
further, a whole vehicle driving force distribution coefficient K is definedd：
Kd＝(Fm+Fr)/(Ff+Fm+Fr) (8)
The practical meaning of the distribution coefficient is the proportion of the driving force of the mechanical transmission path to the total driving force; then, combining the relationship of the rotation speeds of the wheels of the vehicle, the following is obtained:
v＝vf(1-sf)＝vm(1-sm)＝vr(1-sr) (9)
the traction efficiency of a vehicle in all-wheel drive can be further expressed as:
Figure FDA0003068485470000022
according to the formula (10), in the closed hydraulic circuit boosting mode, the system traction efficiency mainly depends on the slip ratio of each wheel and the distribution condition of the whole vehicle driving force in the front wheel hydraulic path and the middle and rear wheel mechanical path of the vehicle;
determining the optimal traction efficiency based on the driving force distribution coefficient; according to the whole vehicle traction efficiency formula determined in the step two, the whole vehicle traction efficiency formula is further subjected to pairing based on the wheel slip rate and the whole vehicle driving force distribution coefficient(10) The derivation is carried out to obtain the system traction efficiency etasTo driving force distribution coefficient KdFirst and second partial derivatives of (c):
Figure FDA0003068485470000023
when the first partial derivative f is expressed by equation (11)1When the driving force distribution coefficient is equal to 0, the corresponding driving force distribution coefficient can enable the traction efficiency of the system to obtain an extreme value, and the slip ratios of the front wheels and the middle and rear wheels have three states: sf1 or sr1 or sf＝sr(ii) a Considering that the wheel slip rate cannot reach 100% in the normal running process of the vehicle, s is equal to and only equal tof＝srIn time, the first-order partial derivative is zero, which is a necessary condition for obtaining the optimal traction efficiency of the whole vehicle; but due to its second partial derivative f2During the change of the slip ratio of the front wheel and the middle and rear wheels, f can not be ensured2If < 0 is always true, s cannot be guaranteedf＝srThe system traction efficiency is maximum, so the extreme point of the formula (11) is further analyzed by combining the change of the driving force distribution coefficient and the states of the speeds of the front wheel and the middle and rear wheels;
firstly, under the static condition that the slip rates of front and rear wheels are not changed; current wheel speed<When the rear wheel is rotating, sf＜srI.e. first partial derivatives f1< 0 is always true; at this time, the distribution coefficient K is varieddThe driving force of a mechanical path is reduced, the driving force of a hydraulic path is increased, and the traction efficiency of the system is gradually increased; and the current wheel speed>When the rear wheel is rotating, sf＞srI.e. first partial derivatives f1If more than 0 is always true; at this time, the distribution coefficient K is varieddThe driving force of a mechanical path is reduced, the driving force of a hydraulic path is increased, and the traction efficiency of the system is gradually reduced; it can be seen that in this static case, when K isd1 or KdWhen the traction efficiency is equal to 0, the system traction efficiency reaches a maximum value;
in the actual running process of the vehicle, the driving force distribution coefficient and the slip ratio of the wheels are not isolated, different driving force distribution ratios have the influence on the change of the slip ratio of the wheels, and the slip ratio of the wheels is increased along with the increase of the driving force, so that the vehicle can not obtain the optimal traction efficiency under the condition that the driving force distribution coefficient is extreme;
therefore, if the front wheel speed is changed during the change of the driving force distribution ratio<Rear wheel speed, with distribution coefficient KdThe driving force of the mechanical path is reduced, the driving force of the hydraulic path is increased, the slip rate of the front wheel is gradually increased, the slip rate of the middle wheel and the rear wheel is also gradually reduced, the traction efficiency of the system is gradually increased, but the corresponding K is obtaineddThe traction efficiency of the system is gradually reduced under the condition of 0; at distribution coefficient KdDuring the reduction process, the wheel speed of the front wheel may exceed that of the rear wheel, and at the moment, the traction efficiency of the system follows KdDecrease to further decrease; it can be seen that with the change of the front and rear wheel speeds, there is always a driving force distribution point to make the traction efficiency of the system reach the maximum value in the dynamic change process, and the point is that the corresponding slip ratio satisfies sf＝srAt the point where the optimum tractive efficiency achievable by the system is ηs,max＝1-sf＝1-sr(ii) a Furthermore, neglecting the difference between the front wheel and the middle and rear wheel, the wheel speed of the front wheel is equal to the wheel speed of the middle and rear wheel;
in summary, when the system works in the closed hydraulic circuit pump boosting mode, the distribution condition of the driving force in the front wheel hydraulic path and the middle and rear wheel mechanical paths should be adjusted as much as possible, so that the wheel speed of the front wheel and the wheel speed of the middle and rear wheels can be followed, and when the control target is reached, the optimal pump displacement target in the mode is as follows:
Figure FDA0003068485470000031
hydraulic variable displacement pump volumetric efficiency etapvAnd the volumetric efficiency eta of the hub hydraulic motormvThe calculation formulas of (A) and (B) are respectively as follows:
Figure FDA0003068485470000032
Figure FDA0003068485470000033
wherein μ represents the dynamic viscosity of the hydraulic oil, CpsRepresents the laminar leakage coefficient, C, of the hydraulic pumpmsRepresenting a laminar leakage coefficient of the hydraulic motor;
and (5) obtaining a final expression form of the target displacement control law of the variable displacement pump according to the expressions (12) to (14):
Figure FDA0003068485470000041
according to equation (15), the target control displacement of the variable displacement pump in the closed hydraulic circuit pump assist mode is composed of two parts: one is the fixed value part associated with the gear,
Figure FDA0003068485470000042
the second is an efficiency compensation part caused by relevant factors such as rotating speed, working pressure and the like,
Figure FDA0003068485470000043
when the vehicle runs at a certain speed, the gear is fixed, and the fixed displacement part beta is usedgOnly depends on the gear of the speed changer and can be obtained by calculation according to the parameters of the whole vehicle, and the efficiency compensation part betaηAnd finally calculating to obtain a pump displacement control target in the closed hydraulic circuit pump boosting mode according to different engine rotating speeds and system pressures related to the system pressure and the engine rotating speed.
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