1. Introduction
Ultra-high precision large stroke displacement is necessary to achieve ultra-precision machining and inspection [
1]. The current high-precision displacement is mainly realized based on intelligent materials’ electrical, magnetic, thermal, optical, and acoustic effects. These have obvious shortcomings, such as tiny stroke, low rigidity, nonlinearity, and hysteresis. Although the ball screw and linear motion (LM) guide can realize large-stroke linear feed, the low-speed creeping phenomenon caused by the inherent properties of its electromechanical system makes it difficult to achieve precise and uniform micro-feeds [
2]. Feng et al. proposed a two-axis differential micro-feed system (TDMS) based on a nut-driven ball screw pair and differential recombination principle [
3]. On the premise of ensuring the micro-feed, the screw speed and nut speed can be increased at the same time. The low-speed nonlinear crawling zone of conventional electromechanical servo systems is avoided. Although TDMS has obtained lower transmission speed and micro-feed than traditional ball screws, it is limited by the high-frequency tremor caused by the machining errors of the balls and raceways and the influence of the backlash of the ball screw. The accuracy of TDMS cannot be further improved. In addition, the high rigidity, high moving accuracy, and almost frictionless transmission characteristic of hydrostatic transmission components can make up for the deficiencies of rolling transmission components. However, they still cannot avoid the low-speed crawling area of the motors [
4].
In order to realize the linear feed with large stroke, low speed, and ultra-high precision, Feng et al. proposed a dual-drive hydrostatic screw micro-nano feed system (DDHLS) based on the nut-driven hydrostatic lead screw [
5]. Since the screw and the nut are supported by a lubricating oil film, DDHLS can significantly reduce the vibration and noise of the TDMS to achieve ultra-precision linear feed. However, in addition to the nonlinear friction of the LM guide, the friction force existing in the nut-driven hydrostatic nut and the shear force of the oil film in the hydrostatic components still interfere with the displacement accuracy of the DDHLS.
For decades, scholars have conducted extensive research on friction modeling and compensation control technology [
6]. Various friction models have appeared successively, such as the Coulomb friction + viscous friction model, Stribeck friction model, Dahl friction model, and Lugre friction model [
7,
8]. Among them, the LuGre friction model can capture the complex static and dynamic characteristics in the friction process, including the Stribeck effect, hysteresis, sticking spring characteristics, and different separation forces. Therefore, the LuGre friction model was selected in this paper. Note that the effect of ultra-low friction is not considered, since the extremely low feed rate of the table in this study exceeds 100 nm/s [
9]. Based on considering torque transmission, Li et al. proposed a dynamic model and parameter identification method to reduce the mechanical tracking error of the ball screw and proposed a feedforward compensation method [
10]. To improve the model estimation performance of the feed system, Lee et al. proposed a fast identification algorithm for feed transmission mass and sliding friction coefficient based on the recursive least squares (RLS) method [
11]. Yang et al. proposed a two-level friction model related to speed and established a tracking error pre-compensation model for CNC machine tools based on the feedforward friction compensation [
12]. Thenozhi et al. identified the parameters of the continuous friction model through a two-step offline identification method and designed a servo-mechanism tracking controller by the backstepping method [
13]. Dumanli et al. improved the tracking accuracy of the machine tool feed through a data-based closed-loop adjustment scheme. The positioning errors caused by servo dynamics and frictional interference were pre-compensated by modifying the reference trajectory [
14]. In order to achieve high-precision tracking control of parallel manipulators, Sancak et al. used a dynamic LuGre model to model joint friction. They also designed the Luenberger-like observer and extended state observer [
15]. Feng et al. described the nonlinear friction of the excavator electro-hydraulic system with the improved Stribeck model and designed a dynamic friction feedforward compensation method based on the structural invariance principle [
16]. In the above study, the movement speed of the actuator can be obtained by multiplying the drive system by the transmission ratio, so it is considered that the drive end and the actuator end have the same friction characteristics.
However, due to the unique transmission structure of DDHLS, different macro-movement speeds can be selected for combination while ensuring the low-speed feed, and the traditional identification methods cannot be used. It should be noted that the friction characteristics of rolling contact components and hydrostatic transmission components are significantly different. The friction torque generated by the liquid friction of DDHLS at different macro speeds also complicates the feed system [
17]. Although Feng et al. realized the separate identification of the ball screw and the LM guide and designed an observer-based PD friction compensator, it no longer applies to the more complex friction conditions of DDHLS [
3]. In addition, the friction torque in the DDHLS is affected by various disturbance factors such as temperature change and mechanical vibration/shock. In order to achieve accurate compensation of friction torque, the friction model of each rolling contact friction component (such as the LM guide and rolling bearing) needs to consider the influence of the external environment [
18]. At the same time, the lubricant viscosity of each hydrostatic component is affected by the heat generation of oil film friction at different speeds. Conventional PID control is challenging to deal with parameter uncertainty, modeling uncertainty, and nonlinearity for high-performance motion control [
19]. Sliding mode control has strong robustness to system parameter perturbations and external disturbances, but the state trajectory is difficult to slide strictly along the sliding surface, which is prone to the chattering phenomenon [
20]. A targeted design of friction compensation controller is required to obtain high control accuracy.
In this paper, the fluid friction model of each hydrostatic component is established, and a calculation method of oil film dynamic friction is proposed based on the variable viscosity theory. Afterward, an all-components refinement friction identification method (ACRFIM) was proposed to model DDHLS precisely. In addition, considering the influence of external disturbance factors, two factors of temperature change and disturbance were introduced to improve the friction model of the rolling contact component. Moreover, considering the influence of oil film dynamic friction, an all-component adaptive friction compensation control algorithm (AACA) is proposed. Finally, the dynamic characteristics and tracking performance of DDHLS at low-speed feed are verified by experiments.
4. Controller Design
In the proposed DDHLS, the motor and the nut drive shaft are highly susceptible to nonlinear friction at low rotational speeds. At the same time, the torque generated by the oil film friction of hydrostatic transmission components varies at different rotational speeds. In addition, the friction model parameters of the rolling contact components are inevitably affected by temperature, wear, lubrication, and position during operation. It is necessary to design the friction compensation controller to achieve high-precision feed control.
Based on the research of Jiang et al. [
18], two influence factors are introduced into the LuGre friction model according to the temperature and disturbance of each rolling contact friction component. The AACA was proposed based on the all-component friction model, as shown in
Figure 9. After that, it was combined with Lyapunov stability theory for stability analysis. When the screw speed is greater than the nut speed, the nut drive motor only needs to offset the friction torque of the nut shaft, and the friction torque is caused by the LM guide acts on the screw shaft. Substituting Equations (14), (19), and (20) into (1) and making
, the following equation can be obtained:
Where is the influence factor of the load force on the bristle offset, and is the change in viscous friction moment due to temperature change. I = m, g, and n represent the screw motor shaft, LM guide, and nut shaft.
Firstly, taking the screw drive shaft as an example, an adaptive friction compensation algorithm with high robustness is designed using the backstepping method. The position tracking error of the screw shaft servo system and the dynamic change of the tracking error are defined as:
where
is the desired rotation angle.
Select the first Lyapunov function and derive it:
The speed control signal is designed as follows:
The integral term is used to ensure that the tracking error of the system approaches zero under the condition of model or load uncertainty. k1 > 0 and k2 > 0 are design parameters.
The second error variable of the screw shaft servo system is designed as follows:
From Equations (19), (20), (27), and (31), the following equation can be obtained:
where
,
,
,
, and
denote the uncertain parameters present in the system, and their real values are represented by the estimated values
,
,
,
, and
. The error between the estimated and actual values can be expressed as:
The average bristle offset
z in the LuGre friction model is unknown and cannot be measured. Here, four nonlinear observers are designed to estimate the
z value. The nonlinear state observer equations are as follows:
where
,
,
,
are the estimated values of
z.
and
are the observer compensation items.
The corresponding state estimation errors and their changes can be expressed as follows:
The Lyapunov function is further defined as follows:
where
k3 > 0 is the design parameter,
r0 > 0,
r1 > 0,
r2 > 0,
r3 > 0, and
r4 > 0 are the adaptive gains. Since
k3,
r0,
r1,
r2,
r3, and
r4 are positive real numbers, it can be determined that
Vs2 is positive definite.
The differentiation of the Lyapunov function with respect to time is calculated by Equations (32) and (37), as follows:
Combining Equations (31) and (32), the control law of the screw axis servo system is designed as follows:
Substitute Equation (39) into Equation (38) and simplify it:
For the screw shaft servo system, the adaptive laws and the observer error compensation terms for each unknown parameter are selected as shown in Equation (41):
The adaptive friction compensation algorithm of the nut drive shaft adopts the same design method as the screw drive shaft, and the control law of the nut shaft servo system is designed as follows:
For the nut shaft servo system, the adaptive laws of the unknown parameters and the observer error compensation term are shown in Equation (43):