Design and Analysis of Thermal Management System of Power Matching Transmission in Energy Machinery

Engineering machinery has aroused widespread attention in the society, as an important part of basic construction equipment and equipment industry. Despite its rapid development, engineering machinery faces several noneligible problems: high energy consumption, low operating efficiency, and excessive heat loss. These problems are inevitable due to the late start of research into engineering machinery. Hence, the energy conservation and emission reduction raise a growing concern in the industry [1-4].

The installation of auxiliary energy-saving devices increases the internal density of engineering machinery. The rising full-machine heat load that ensues challenges the effect of the cooling system [5-8]. To enhance the full-machine performance in dynamics, economy, and greenness, it is important to carry out energy analysis, power matching, and full-machine thermal management of engineering machinery, especially loader.

Under the premise of ensuring the speed control and overload capacity of engineering machinery, replacement engine with motor is an effective way to realize an energy-efficient, low-emission drive mode [9-10]. To improve the energy efficiency, operability, dynamics, and safety of the full machine, it is of great significance to study the power matching structure and control strategy of engineering machinery under complex conditions [11-14]. Wang et al. [15] analyzed the load features and working conditions of the hydraulic excavator driven by permanent magnet synchronous motor, put forward a power assembly parameter matching method that improves the full-machine performance, and provided a variable pressure difference control strategy suitable for low-speed small flows. To improve the performance of wheel loader transmission system, Niu et al. [16] constructed a dynamic simulation model of loader transmission system to reduce fuel consumption in each cycle and heat generation, optimized the main parameters of loader transmission by genetic algorithm (GA), and examined the influence of different flow modes on the flow field inside the oil cooler, in view of the cooling demand of the transmission system. Using the performance calculation method based on heat transfer units (HTUs), Wang et al. [17] investigated the relationship between the surface air flow rate and air-side resistance of the radiator of engineering machinery, clarified the method for cooling fan selection and the design steps of variable circulation, and designed a full-machine thermal balance test, revealing that the prototype with two circulating cooling systems has relatively low energy consumption. To prevent the overheating and overcooling of engineering machinery, Nguyen et al. [18] selected the main components and discussed the control strategy for the electronically controlled hydraulic fan cooling system based on BODAS controller, and evaluated the energy efficiency of the cooling system by measuring and comparing the heat release of the system and the economy of the engine.

Heat dissipation is needed for both the engine of engineering machinery and the hydraulic oil of hydraulic torque converter. The huge cooling demand cannot be satisfied with one cooling fan alone [19-22]. Naumenko et al. [23] carried out the matching design for the cooling fan and radiator of engineering machinery, reconfigured the thermal parameters and component parameters of the hydraulic drive and transmission, and conducted experiments to compare the engine water temperatures and preheating times before and after system improvement; their electro-hydraulic hybrid drive cooling system was proved effective in cooling.

In the traditional thermal management system for the transmission in energy machinery, the cooling capacity does not fully match the load of each subsystem, and the thermal management components cannot adapt to the dynamic cooling demand of engineering machinery. To solve these problems, this paper designs and analyzes the thermal management system of power matching transmission in energy machinery. The target engineering machinery is a loader. Section 2 analyzes the energy distribution among different components of engineering machinery, including, hydraulic system, transmission system, and cooling system in each work section, and provides a dynamic matching method for hydraulic torque converter. Section 3 determines the targets of thermal management for the power transmission system of loader: heat source engine, hydraulic torque converter, and hydraulic retarder, prepares as a scheme in which the engine transmission system is independent with the cooling circulation system, takes the scheme as the framework of the thermal management cycle, and constructs a heat transfer calculation model for loader transmission system. The proposed model was proved effective through experiments.

The thermal management of loader transmission has three targets: heat source engine, hydraulic torque converter, and hydraulic retarder. There are two schemes for thermal management cycle. In the first scheme, the engine transmission and cooling circulation systems are coupled; in the second schemes, the two systems are independent of each other. On the upside, the first scheme boasts a compact thermal management system, and low efficiency of secondary heat transfer; on the downside, the system is too complicated to be maintained at a low cost. The second scheme features a simple structure, reliable operations, and a high cooling efficiency. Besides, the weak coupling between control variables makes it easy to implement the control strategy. To rationalize the control system of thermal management system, this paper decides to study the dynamic transmission system and thermal management system separately, i.e., adopt the second scheme to analyze the thermal load and thermal performance of the two systems (Figure 3).

3.png

Figure 3. Working principle of Scheme 2

This section aims to model the structure of the thermal management system for loader engine, which consists of water jacket, water pump, thermostat, radiator, expansion tank, and heater. In the thermal management system, the power and temperature cycle of the engine are respectively controlled by water pump and thermostat. The radiator is responsible for the heat exchange between the engine and the environment after the start of the cooling cycle. The pressure stability of the system is mainly achieved by bubble discharge of the expansion tank. Let S_CL be the equivalent cross-sectional area of the cooling pipe; ∆F be the inlet-outlet pressure difference of the coolant of the heat management component; φ_C be the density of the coolant. Then, the volumetric flow L_V of the engine’s heat management component can be described by modified Bernoulli’s equation:

$L_{V}=S_{C L} \cdot \sqrt{\frac{2 \cdot \Delta F}{\varphi_{C}}}$    (12)

Let E_TAH be the heat exchange amount between the engine housing and the cooling pipe; υ_M and ω_M be the engine speed and the corresponding maximum power. Then, the heat exchange amount between the engine and the coolant can be calculated by:

$E_{T A H}=g\left(v_{M}, \omega_{M}\right)$    (13)

Let σ_M and EU be the heat capacity unit of the cooling pipe and the resistance unit of the system, respectively; F_M and Tψ_M be the pressure and temperature of the hot fluid passing through the heat capacity unit, respectively; D_M be the volume of the cooling pipe of the engine; dE be the heat entering that volume; σ_F be the specific heat of the fluid, which depends on the fluid’s working pressure and temperature; φ_C be the fluid density. Then, the differentials of T_M and F_M relative to time, denoted respectively as dT_M/dt and dF_M/dt, must satisfy:

$\frac{d T_{M}}{d t}=\frac{d E}{\sigma_{F} \cdot \varphi_{C} \cdot D_{M}}+\frac{d T_{M}}{\varphi_{C} \sigma_{F}} \cdot \frac{d F_{M}}{d t}$    (14)

Let dqe₁, dqe₂ and dq₁, dq₂ be the heat flow and mass flow entering the heat capacity unit of the cooling pipe; de be the heat exchange amount between the heat capacity unit and the environment; e be the heat of the fluid in the unit. Then, dE can be calculated by:

$d E=d q e_{1}+d q e_{2}+d e-\left(d q_{1}+d q_{2}\right) \cdot e$    (15)

Let dφ_Fbe the variation of density; δ and β be the elastic modulus and volume expansion coefficient of the liquid, respectively. Both δ and β depends on the working pressure and temperature of the medium. Then, the pressure per unit volume can be described by the following differential equation:

$\frac{d F_{M}}{d t}=\delta \cdot \frac{d \varphi_{E}}{\varphi_{E}}+\delta \cdot \beta \cdot \frac{d T_{M}}{d t}$    (16)

The inlet/outlet mass flow dq' and heat flow dq'·e' of the resistance unit of the engine can be derived based on the resistance unit of the cooling system. Assuming that the inlet and outlet heat flows are equal for each resistance unit, dq' can be calculated as by:

$d q^{\prime} e_{1}^{\prime}=d q^{\prime} \cdot e_{1}^{\prime}$    (17)

$d q^{\prime} e_{2}^{\prime}=d q^{\prime} \cdot e_{1}^{\prime}$    (18)

The cooling system of the engine is powered by the centrifugal water pump driven at a fixed speed by the crankshaft of the engine. Let υ_M be the speed of the engine; i_P/M be the speed ratio of the cooling water pump to the engine. Then, the speed of the cooling water pump can be calculated by:

$v_{W P}=v_{M} \cdot i_{P / M}$    (19)

The displacement and speed of the cooling water pump determine its inlet-output pressure difference. The fluid temperature at the outlet of the water pump can be obtained from the measured temperature of the inlet fluid and the calculation results on the pump’s work on the fluid. Here, the cooling water is treated as an ideal incompressible fluid. Let γ_a be the convection coefficient between coolant and oil; σ_W be the heat capacity of the oil; T_EC be the coolant temperature at the inlet of the thermostat. Then, the power of the water pump can be calculated by:

$\frac{d T_{W P}}{d t}=\frac{\gamma_{a}}{\sigma_{W}} \cdot\left(T_{E C}-T_{W P}\right)$    (20)

The flow regime of coolant depends on the Reynolds number. For the water-air radiator, the volumetric flow of the coolant can be determined by the inlet and outlet pressures, and used to deduce the rated heat exchange amount of the radiator. The heat exchange amount between the radiator and the coolant can be characterized by the function between the volumetric flow of coolant and the mass flow of the air. The relationship between the inlet temperature T_EHS, outlet temperature T_OHS, and inlet-outlet temperature difference dT_HS of the radiator can be described as:

$d T_{H S}=T_{E H S}-T_{O H S}$    (21)

Let ψ_Eact, ψ_Oact, and E_HS be the measured inlet temperature, measured outlet temperature, and heat release of the radiator, respectively; P_HS be the thermal efficiency of the radiator. Then, the actual heat release E_act of the radiator can be calculated by:

$E_{a c t}=E_{H S} \cdot\left(T_{\text {Eact }}-T_{\text {Oact }}\right) / d T_{H S} \cdot P_{H S}$    (22)

The cooling effect of the radiator differs significantly between ventilation area and non-ventilation area. To calculate the total heat exchange amount of the radiator, it is necessary to compute the heat exchange amounts of the two areas, both of which are related to air flow rate. Let υ_NT, and υ_AFR be the air flow rates before and after the radiator fans are opened; υ_OA be the additional wind speed brought by the fan. Then, we have:

$v_{A F R}=v_{N T}+v_{O A}$    (23)

Ignoring the air flow at the center of the radiator fans, the overall mapping area S_FMA of the fan can be obtained by adding up the actual coverage of the fan with the rotational area at the center of the fan. Let ON_fan and IN_fan the outer and inner diameters of the fan, respectively. Then, the S_FMA value can be calculated by:

$S_{F M A}=\frac{\pi}{4}\left(O N_{f a n}^{2}-I N_{f a n}^{2}\right)$    (24)

Let Ф and U be the length and height of the radiator, respectively. Then, the non-mapping area of the fan S_NFMA can be calculated by:

$S_{N F M A}=\Phi \cdot U-S_{F M A}$    (25)

Suppose the heat exchange amount of the radiator is determined by the air flow rate in the mapping area and the volumetric flow of the coolant. The functional relationship between the two determinants can be illustrated as g(υ_FMA, L_R). Then, the heat exchange amount E_FMA within S_FMA can be expressed as:

$E_{F M A}=g\left(v_{F M A} \cdot L_{R}\right) \cdot \frac{S_{F M A}}{\Phi \cdot U}$    (26)

Suppose the heat exchange amount of the radiator is determined by the air flow rate in the non-mapping area and the volumetric flow of the coolant. The functional relationship between the two determinants can be illustrated as g(υ_NT, L_R). Then, the heat exchange amount E_NFMA within S_NFMA can be expressed as:

$E_{N F M A}=g\left(v_{N T} \cdot L_{R}\right) \cdot \frac{S_{N F M A}}{\Phi \cdot U}$    (27)

The total heat exchange amount E_T can be obtained by:

$E_{T}=E_{F M A}+E_{N F M A}$    (28)

Let φ_F and SN_G be the density of the working medium and the effective diameter of the driving wheel of hydraulic retarder, respectively; a be the gravity acceleration; μ_TC and AV_G be the braking torque coefficient and the annular velocity of the drive wheel, respectively. Then, the braking torque of hydraulic retarder can be calculated by:

$T_{C}=\varphi_{E} \cdot a \cdot \mu_{T_{C}} \cdot A V_{G}^{2} \cdot S N_{G}^{5}$    (29)

Let Q_LP and M_SL be the heat flow of power loss and braking torque of the retarder, respectively; r_M be the engine speed. Since the transmission fluid in the retarder contains much more kinetic energy than the fluid, Q_LP can be empirically approximated by:

$Q_{L P}=\frac{M_{S L} \cdot r_{M}}{9550}$    (30)

The cyclic flow features of the retarder can be characterized by the effective displacement of the variable pump. Let υ_ED and υ_FD be the effective displacement and displacement baseline of the variable pump, respectively; η_VPD be the proportional coefficient of the variable displacement of the variable pump. Then, the υ_ED value can be calculated by:

$v_{E D}=v_{F D} \cdot \eta_{V P D}$    (31)

Let L_VP and r_VP be the volumetric flow and speed of the variable pump, respectively. Then, the rated volumetric flow of the pump can be calculated by:

$L_{V P}=v_{E D} \cdot r_{V P}$    (32)

Under different working speeds, the retarder sees changes in transmission fluid flow, inlet pressure, and output pressure. Any change to the flow in the thermal management system will change the pressure increment of the retarder. Therefore, the proposed thermal management system for the full-machine power transmission system must simulate the correspondence between these changing parameters with a pilot pump. Let f_s be the lever position of the retarder. Then, the relationship between the outlet and inlet pressures of the retarder can be defined as:

$F_{O}=F_{I}+f_{s}$    (33)

This paper designs and analyzes the thermal management system of power matching transmission in energy machinery. Firstly, the authors described the energy distribution among different components of engineering machinery, including, hydraulic system, transmission system, and cooling system in each work section, and detailed the dynamic matching method for hydraulic torque converter. After that, the dynamic features of the torque converter were tested, and the engine speed curve throughout the driving process was plotted based on the test results. Then, heat source engine, hydraulic torque converter, and hydraulic retarder were selected as the targets of thermal management for the power transmission system of loader, the scheme in which the engine transmission system is independent with the cooling circulation system was adopted as the framework of the thermal management cycle, and the corresponding heat transfer model was built for the loader transmission. During the experiments, the fans were controlled by two different methods: traditional mechanical control with fixed speed ratio, and temperature control. Under the two control modes, the authors investigated the variations in the coolant temperature of engine, mean parasitic power of engine, transmission fluid temperature of radiator, and the power consumption of radiator fans. The experimental results confirm the effectiveness of our model, and provide a reference for the thermal management of other subsystems of engineering machinery.