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Automotive Innovation

, Volume 2, Issue 1, pp 35–44 | Cite as

Comparative Study on the Temperature Rise of a Dry Dual Clutch Under Different Starting Conditions

  • Maotao Zhu
  • Peng YaoEmail author
  • Yubin Pu
  • Tao Liu
Article
  • 96 Downloads

Abstract

A finite element model of the pressure plate and friction plate of the dry dual clutch is established to find the temperature rise of a dry double clutch in operation. Different starting conditions for the temperature of the dry dual clutch are obtained by calculating the temperature rise, and a comparative analysis of the results is conducted. Results show that the temperature rise for clutch 2 is generally higher than that for clutch 1, and the temperature rise of clutch 2 is higher than that of clutch 1, with an increasing throttle opening. A curve of the speed difference between engine and friction plates is extracted by calculating the source power of the wear of the sliding mill. It is found that the clutch speed difference of clutch 2 is generally higher than that of clutch 1 because clutch 1 follows a constant-rotation-speed control strategy at the start of grinding, leading to the temperature rise of clutch 2 being greater than that of clutch 1. A constant-engine-speed shift strategy is finally put forward.

Keywords

Dry double clutch Friction power Temperature field Comparative analysis Speed difference 

1 Introduction

The dry dual-clutch automatic transmission not only has the advantages of high efficiency, compactness, and low weight, but also reduces the vehicle acceleration time and fuel consumption [1]. However, owing to the low heat capacity of the dry dual clutch, it is not easy to distribute heat generated by the clutch in a short time, especially in cases of urban traffic congestion. The clutch temperature increases rapidly owing to repeated combination, which directly increases friction wear [2, 3]. The dry dual-clutch assembly is riveted in assembly, and the whole assembly thus has to be replaced if any part of the clutch (i.e., the pressure plate or friction plate) wears excessively. The life of the dry dual-clutch assembly is therefore decided by the part experiencing a severe temperature rise [4]. A comparative analysis of the temperature rises of the two clutches under different starting conditions is therefore important for improving the life of the dual clutch.

Xu et al. [2] proposed that the node temperature rises with an increase in the radius of the pressure plate by analyzing one-third of the pressure plate. Liu et al. [5] analyzed the effect of the heat capacity of the pressure plate in the starting process and found that the maximum temperature can be reduced by 40% by doubling the heat capacity. Schwab analyzed strong effects of clutch combination characteristics in terms of the temperature rise during the shift process [6]. Wu and Si [7] analyzed the effect of a control strategy on the temperature rise of the starting process of the dual clutch and put forward an intelligent control strategy. In summary, the above research mainly focused on the temperature rise of the pressure plate and the control strategy, and there has been no comparison of the temperature rise between the two types of clutch under different starting conditions. It is thus not known what part of the dual clutch is severely affected under different starting conditions.

The present paper establishes a vehicle dynamics model using the simulation software AVL Cruise, and calculations are made for three types of starting conditions (i.e., a slow start, normal start, and quick start) to attain the friction work and power of the dual clutch, which are used to calculate the heat flux as the boundary of the temperature calculation. A computational fluid dynamics (CFD) simulation is conducted to analyze the surrounding air movement of the pressure plate and the friction plate under eight speeds, and the convective heat transfer coefficient is calculated and used to calculate the temperature field. The transient temperature field of the pressure plate and friction plate is analyzed, and the results of temperature rise are compared and analyzed. The reason why the sliding temperature rise of clutch 2 is generally higher than that of clutch 1 is explained on the basis of the angle of the rotating-speed difference between the engine and clutch. The paper finally puts forward a control strategy that minimizes the speed difference between clutch 2 and the engine and reduces the temperature rise of clutch 2.

2 Vehicle Dynamics Simulation

2.1 Dynamic Model of the Dual-Clutch Transmission (DCT)

It is difficult to establish an accurate dynamic model because the vehicle drive system is a continuous and complex system with multiple degrees of freedom. The following assumptions are made to improve the computational efficiency [5].
  1. (1)

    Lateral vibration of the shaft can be ignored.

     
  2. (2)

    The effects of the system gap and damping can be ignored.

     
  3. (3)

    The effect of the elasticity of the bearing and the bearing housing can be ignored.

     
On the basis of the above assumptions, a simplified DCT dynamics model is presented in Fig. 1.
Fig. 1

Dynamic model of the DCT

2.2 Engine Model

Figure 2 presents a characteristic plot of engine torque obtained by experiment.
Fig. 2

Characteristic plot of the engine torque

The figure shows that the output torque is related to the throttle opening and engine speed [8]:
$$T_{\text{e}} = f\left( {\beta ,n_{\text{e}} } \right) $$
(1)
where Te is the output torque of the engine under a steady condition, N m; ne is the engine speed, r/min; and β is the throttle opening, ranging from 0 to 1.
As the vehicle is usually in an unstable condition (in terms of the shift, acceleration, and deceleration), it is important to establish an (unsteady) engine dynamic model for DCT dynamic simulation [3]:
$$T_{\text{ed}} = T_{\text{e}} - \alpha \varpi_{\text{e}} $$
(2)
where Ted is the output torque of the engine under a dynamic condition (i.e., non-steady state), N m; \(\varpi_{\text{e}}\) is the angular acceleration of the engine crankshaft, rad/s2; and α is the falling coefficient of the engine output torque under dynamic conditions (i.e., non-steady state), which depends on the type of engine.

2.3 Model of the DCT

2.3.1 Starting Model of the DCT

The present paper adopts a single-clutch starting model, in which only clutch 1 takes part in sliding when the vehicle starts. The start process is divided into sliding and engagement stages [2]. Therefore, a mathematical model of the powertrain system at the start is first established:
$$T_{\text{e}} - T_{{{\text{c}}1}} = J_{\text{e}} \dot{w}_{\text{e}} $$
(3)
$$T_{{{\text{c}}1}} i_{1} - T_{\text{n}} /i_{0} = J\dot{w}_{\text{v}} $$
(4)
$$T_{\text{n}} = \left[ {(f\cos \beta + \sin \beta )mg + \frac{{C_{\text{D}} A(3.6w_{\text{v}} r_{\text{r}} )^{2} }}{21.15}} \right]r_{\text{r}} $$
(5)
$$J = J_{{{\text{c}}1}} i_{1}^{2} + mr_{\text{r}}^{2} /i_{0}^{2} $$
(6)
where Te and Tc1 are, respectively, the output torque of the engine and transmission torque of clutch 1, N m; Tn is the torque of the driving resistance, N m; we and wv are, respectively, the speeds of the engine and transmission output shaft, r/min; Je is the rotational inertia of the rotating part of the engine and the active part of the clutch, kg m2; Jc1 is the rotational inertia of the driven part and connecting part of clutch 1, kg m2; i1 is the transmission ratio of the first gear of the gearbox; i0 is the main drive ratio of the vehicle; rr is the tire rolling radius, m; f is the rolling resistance coefficient; β is the road grade; m is the cross weight, kg; g is acceleration due to gravity, m/s2; CD is the air resistance coefficient; and A is the frontal area, m2.
  1. (1)

    Sliding Stage

     
In the sliding stage, the gap between clutch 1 and the friction plate is eliminated and the platen and friction plate begin to slide against each other, while clutch 2 remains separated. The frictional torque of clutch 2 therefore remains zero, while the output torque of clutch 1 is
$$T_{{{\text{c}}1}} = \mu_{1} F_{1} Z_{1} \times \frac{2}{3}\left( {\frac{{R_{10}^{3} - R_{11}^{3} }}{{R_{10}^{2} - R_{11}^{2} }}} \right)$$
(7)
where μ1 is the frictional coefficient of clutch 1, which is a variable parameter affected by the angular velocity difference between the driving plate and driven plate, the pressing force of the friction plate, and the temperature during the clutch engagement and separation process; F1 is the pressing force of clutch 1, N; Z1 is the number of friction surfaces of clutch 1, Z1 = 2; and R10 and R11 are, respectively, the inner and outer radii of the friction surfaces of clutch 1, mm.
  1. (2)

    Engagement Stage

     
In the engagement stage, clutch 1 ceases sliding and remains in a state of engagement, and the transmitted torque is not equal to the output torque of the engine. Tc1 is derived from \(w_{\text{e}} = i_{1} w_{\text{v}}\) and Eqs. (3) and (4):
$$T_{{{\text{c}}1}} = (JT_{\text{e}} + i_{1} T_{\text{n}} J_{\text{e}} /i_{0} )/(J + i_{1}^{2} J_{\text{e}} )$$
(8)

The constant-speed control strategy not only keeps the engine at a lower speed to reduce the sliding power during launch but also meets the load on the torque requirements. The strategy is thus used in the starting process.

2.3.2 Shifting Model of the DCT

As the vehicle begins moving, clutch 2 is still in the uncombined state and thus goes through four stages: pre-upshift stage → torque stage → inertia stage → high-gear stage [7]. The dynamics model of the two clutches transmitting the torque simultaneously is
$$T_{\text{e}} - (T_{{{\text{c}}1}} + T_{{{\text{c}}2}} ) = J_{\text{e}} \dot{w}_{\text{e}}$$
(9)
$$T_{{{\text{c}}1}} i_{1} + T_{{{\text{c}}2}} i_{2} - T_{\text{n}} /i_{0} = J^{\prime } \dot{w}_{\text{v}}$$
(10)
$$J^{\prime } = J_{{{\text{c}}1}} i_{1}^{2} + J_{{{\text{c}}2}} i_{2}^{2} + mr_{\text{r}}^{2} /i_{0}^{2}$$
(11)
where Tc2 is the transmission torque of clutch 2, N m; i2 is the transmission ratio of the second gear of the gearbox; and Jc2 is the rotational inertia of the driven and connecting parts of clutch 2, kg m2.
  1. (1)

    Pre-upshift Stage

     
In the pre-upshift stage, clutch 1 has completed engagement, while clutch 2 is in a separated state, but the gap between the pressure plate and friction plate of clutch 2 has been eliminated. The plates then engage, and the output torque of the clutch and transmission is consistent with that in the engagement stage.
  1. (2)

    Torque Stage

     
In the torque stage, clutch 1 starts to separate. Clutch 1 is still engaged, but the transmitted frictional torque gradually reduces, and clutch 2 starts to engage. The corresponding output torque is
$$T_{{{\text{c}}2}} = \mu_{2} F_{2} Z_{2} \times \frac{2}{3}\left( {\frac{{R_{20}^{3} - R_{21}^{3} }}{{R_{20}^{2} - R_{21}^{2} }}} \right)$$
(12)
$$T_{{{\text{c}}1}} = \left[ {J^{\prime } (T_{\text{e}} - T_{{{\text{c}}2}} ) - i_{1} i_{2} J_{\text{e}} T_{{{\text{c}}2}} + i_{1} T_{\text{n}} J_{\text{e}} /i_{0} } \right]/(J^{\prime } + i_{1}^{2} J_{\text{e}} )$$
(13)
where F2 is the pressing force of clutch 2, N; μ2 is the frictional coefficient of clutch 2; Z2 is the number of friction surfaces of clutch 2, Z2 = 2; and R20 and R21 are, respectively, the inner and outer radii of the friction surfaces of clutch 2, mm.
  1. (3)

    Inertial Stage

     
In the inertial stage, clutch 1 continues to separate and begins to slide and the torque transmitted by clutch 2 continues to increase, while the output torque of the engine should change accordingly such that clutch 2 rotates synchronously with the flywheel:
$$T_{{{\text{c}}1}} = \mu_{1} F_{1} Z_{1} \times \frac{2}{3}\left( {\frac{{R_{10}^{3} - R_{11}^{3} }}{{R_{10}^{2} - R_{11}^{2} }}} \right)$$
(14)
$$T_{{{\text{c}}2}} = \mu_{2} F_{2} Z_{2} \times \frac{2}{3}\left( {\frac{{R_{20}^{3} - R_{21}^{3} }}{{R_{20}^{2} - R_{21}^{2} }}} \right) .$$
(15)
  1. (4)

    High-Gear Stage

     
In the high-gear stage, clutch 1 has disengaged, while clutch 2 remains engaged:
$$T_{{{\text{c}}2}} = (J^{\prime } T_{\text{e}} + i_{2} T_{\text{n}} J_{\text{e}} /i_{0} )/(J^{\prime } + i_{2}^{2} J_{\text{e}} ) $$
(16)

The present paper analyzes the friction temperature rise in the first and second processes only, so do not repeat the downshift process.

The two-parameter shift schedule with the speed and throttle opening taken as parameters is widely used for vehicles. Meanwhile, the combined schedule provides economy for a small throttle opening and dynamic performance for a large throttle opening. The present paper thus uses this schedule for the first and second processes [7].

The engine dynamic model, dry DCT, and driving resistance subsystem are modeled in the software AVL Cruise [8, 9], as shown in Fig. 3. A slow start (10% throttle opening), normal start (25% throttle opening), and fast start (45% throttle opening) are simulated. To ensure comparability, the speeds of engagement of the two clutches are set the same, and the throttle opening for the same simulation process remains unchanged under different starting conditions, engine speed ωe, clutch speeds ωc1 and ωc2, friction work Wf, and friction power Pf, as shown in Figs. 4, 5, 6, 7, 8, and 9.
Fig. 3

Vehicle dynamics model

Fig. 4

Speeds of the engine and clutch in a slow start

Fig. 5

Friction work and power of the clutch in a slow start

Fig. 6

Speeds of the engine and clutch in a normal start

Fig. 7

Friction work and power of the clutch in a normal start

Fig. 8

Speeds of the engine and clutch in a fast start

Fig. 9

Friction work and power of the clutch in a fast start

The results show that the friction work of clutch 1 depends slightly on the start condition, being 4500, 5500, and 6200 J for slow, normal, and fast starts, respectively. The friction power rises with the throttle opening, being 9000, 12,000, and 14,500 W for slow, normal, and fast starts, respectively. The friction work of clutch 2 is 5100, 6500, and 12,000 J for slow, normal, and fast starts, respectively, and the rise of friction power is faster than that for clutch 1, respectively, 11,000, 18,000, and 26,000 W. It is seen that under the same starting condition, the friction work and power of clutch 2 are higher than those of clutch 1 and are much higher for a fast start than for other starting conditions. The differences between the two clutches are explained in Sect. 5.

3 CFD Simulation Analysis of the Dual Clutch

The heat generated by sliding friction between the pressure plate and friction plate is mainly carried away by heat convection with the surrounding air, and the heat dissipation effect can reach about 80% of the total heat dissipation in the process of the engagement of the clutch [5]. According to heat transfer theory, thermal convection is determined by the motion of air around the pressure plate and friction plate [10, 11]. It is thus important to calculate the air motion around the pressure plate and friction plate in the process of grinding.

This section conducts CFD simulation of the pressure plate and friction plate under eight speed conditions (i.e., 500, 1000, 1500, 2000, 2500, 3000, 3500, and 4000 r/min) [12]. The air flow around the pressure plate is similar at different speeds, with the only difference being the wind speed, and the air flow result is therefore analyzed for a speed of 2000 r/min only. The calculation results of the wind speed around the pressure plate and friction plate are presented in Fig. 10.
Fig. 10

Calculation results of the wind speed at 2000 r/min

Figure 10a, c shows that, because the linear speeds of the outer diameters of the pressure plate and friction plate are greater than those of the inner diameters during rotation, the lateral air flow rate is higher (as indicated by the yellow–green color). At the same time, the air flow on the outside of the pressure plate is disturbed on account of the convex lug, and the flow velocity around the convex lug is obviously higher than that at other positions on the pressure plate. In addition, the air contact surface of the pressure plate of clutch 1 is flat, such that the air flow rate is slightly higher, while the air contact surface of the pressure plate of clutch 2 has a support platform in the inner diameter, which increases the air flow on the inner side (as indicated by the green color). At the same time, owing to structural differences between the two clutches, the thickness of the air layer in the process of rotation is different, resulting in a low air flow rate at the air contact surface of the pressure plate of clutch 2 (as indicated by the deep blue color).

Figure 10b, d shows the flow fields calculated for the friction plates. The surface of the friction plate rotates to the outer diameter of the friction plate owing to the staggered grooves on the sliding surface. However, the rivets between the grooves affect the air flow, such that the air flow rate is low. The flow rate is higher on the inner diameter and the outer diameter of the friction plate, where there is no obstruction. In summary, the flow velocities across the pressure plate and friction plate and the air contact surface depend on the position. The positions of different features are illustrated in Fig. 11. The heat transfer coefficient h is presented in Fig. 12:
$$h = 0.0296\frac{\lambda }{L}R_{\text{e}}^{4/5} P_{\text{r}}^{1/3} $$
(17)
where λ is the thermal conductivity of air, W/(m2 K); L is the feature size, m; Re is the Reynolds number; and Pr is the Prandtl number.
Fig. 11

Schematic diagram of different feature positions

Fig. 12

Calculation results of the heat transfer coefficient

4 Temperature Analysis of the Dual Clutch

4.1 Determination of Heat Flux

In the process of starting the motion of the vehicle and shifting gear, the pressure plate and friction plate move with circular rotary motion (Fig. 13). In the process of grinding, the velocity of the moving line increases with the radius of the pressure plate and more heat is generated. The heat flux of the friction surface is therefore derived from the sliding friction power referring to the previous literature [10]:
$$q(t) = \frac{{{\text{d}}N}}{{{\text{d}}S}} = \frac{3r}{{2\pi \left( {r_{2}^{3} - r_{1}^{3} } \right)}}N(t)$$
(18)
where N(t) is the friction power, W; r1 and r2 are, respectively, the inner and outer radii of the pressure plate, m; and r is the radius of the studied point, m.
Fig. 13

Schematic of the analysis of the generated friction heat

The heat flux density on the sliding surfaces of the pressure plate and friction plate can be calculated according to the friction power and Eqs. (18) and (19).

4.2 Determination of the Heat Distribution Factor

Because of the different properties of the materials of the pressure plate and friction sheet, the amounts of heat transferred to the pressure plate and friction plate in unit time are different; K is expressed as [13, 14]:
$$K = \sqrt{ {\frac{{\lambda_{1} c_{1} \rho_{1} }}{{\lambda_{2} c_{2} \rho_{2} }}}}$$
(19)
where λ1, λ2, c1, c2, ρ1, and ρ2 are, respectively, the thermal conductivity, specific heat capacity, and density of the friction plate and pressure plate (Table 1).
Table 1

Material properties of the friction and pressure plates

Parameters

Friction plate

Pressure plate

λ (W/m2 °C)

1.4

48.2

c (J/kg °C)

1400

470

ρ (kg/m3)

1940

7120

4.3 Application of the Finite Element Boundary Condition

Setting an initial temperature of 25 °C, the calculated heat flux is applied to the sliding surfaces of the pressure and friction plates using the APDL language in a two-dimensional table, which gives the time-dependent heat flux and radius. The heat transfer coefficients of the pressure plate and friction plate, which are calculated in Sect. 2, are included in the CFD simulation analysis.

4.4 Analysis of the Temperature Field of the Dual Clutch

Only the results of the temperature field under the quick start condition are presented, as shown in Figs. 14, 15, 16, 17, 18, and 19, because the temperature distribution and trend of the clutch pressure plate and friction plate are basically the same under different starting conditions.
Fig. 14

Odd platen peak-temperature nephogram for a fast start

Fig. 15

Odd platen temperature nephogram at 2.5 s for a fast start

Fig. 16

Odd friction plate nephogram for a fast start

Fig. 17

Even platen peak-temperature nephogram for a fast start

Fig. 18

Even platen temperature nephogram at 2.5 s for a fast start

Fig. 19

Even friction plate nephogram for a fast start

The temperature nephograms in Figs. 14, 15, 16, 17, 18, and 19 show that the temperature distribution is not uniform at the pressure and friction plates, with the temperature gradient being large. The temperature is maximum at the edge of the maximum friction radius, which is consistent with the expression of heat flux density (18). Because the surface of the pressure plate is far from the sliding surface and the grinding time is short, there is almost no temperature change at the surface when it reaches the peak value. The peak temperatures of the two clutch pressure plates are basically the same as those of the friction plates when the pressure and friction plates are in contact with each other during the grinding process. The temperatures of the pressure and friction plates should therefore be equal, which demonstrates the correctness of the simulation.

Figures 15 and 18 show the temperatures of the two clutches at 2.5 s for a fast start. It is seen that the temperature of the sliding surface of the pressure plate is lowered by convective heat transfer. However, there is still a temperature rise of 5–10 °C, which shows that the convective heat transfer is not able to completely eliminate the heat generated by the sliding mill. In the case of urban road congestion, therefore, owing to the repeated starting and stopping of the vehicle, the clutch temperature can rise rapidly.

5 Comparative Analysis

5.1 Comparison of Results

The peak temperatures of the two types of clutch under the three starting conditions are arranged in Table 2.
Table 2

Summary of peak temperatures

Condition

Clutch 1 (°C)

Clutch 2 (°C)

Pressure plate

Friction plate

Pressure plate

Friction plate

Slow start

33.788

34.421

40.569

41.579

Normal start

36.831

37.865

45.786

47.406

Fast start

41.786

43.025

58.207

60.442

The table shows that the peak temperatures of the pressure and friction plates of clutch 2 are higher than those of clutch 1 under different starting conditions; i.e., the temperature rise is higher for clutch 2. The temperature rise of clutch 2 is thus more serious, and with an increase in the throttle opening, the temperature rise will increase correspondingly, especially under the fast-start condition.

5.2 Reasoning

The result that the temperature of clutch 2 is higher than that of clutch 1 is analyzed by investigating the causes of the clutch temperature rise. The root of the temperature rise is slipping, and the formula for calculating the friction power during slipping is thus combined and simplified as
$$P_{\text{f}} = T \cdot \Delta \omega = T \cdot \left( {\omega_{\text{e}} - \omega_{\text{c}} } \right)$$
(20)
where T is the friction torque of the clutch, N m; ωe is the engine speed, r/min; and ωc is the clutch speed, r/min.
Equation (20) reveals that the friction power is only related to the friction torque and the rotational speed difference, owing to the binding speed of the clutch being the same under different starting conditions. The difference in the sliding power of the two types of clutch is therefore mainly due to the difference in the speed between the engine and the clutch friction plate. Figure 20 presents speed difference curves under different starting conditions.
Fig. 20

Speed differences under different loads

Figure 20 shows the engine speed, friction disk speed, and speed difference curve for a slow start, normal start, and fast start. The red dashed line showing the speed difference between the engine and friction plate 1 and the blue dashed line showing the difference between the engine and friction plate 2 are obviously different. The red dashed line is nonzero and gradually declines in the sliding stage of clutch 1. The blue dashed line falls rapidly from the black curve showing the engine speed, after the first turning point to rise and then fall. The blue dashed curve falls to the first turning point in the pre-engagement stage of the second gear, and this is followed by the grinding stage of clutch 2.

The differences between the engine speed and the speeds of clutches 1 and 2 in each sliding stage are analyzed for the same start condition. It is seen that the engine speed increases slightly in a short time after clutch 1 begins to slip, and the engine speed remains constant until the end of slipping. However, the engine speed continues to increase during the sliding stage of clutch 2. This is because the engine follows a constant-speed control strategy in the grinding stage of clutch 1, which reduces the speed difference to zero. Meanwhile, clutch 2 not only fails to reduce the speed difference owing to the lack of a corresponding speed control strategy but also has a rising trend for a period of time after the start of friction, which leads to clutch 2 having a higher speed difference than clutch 1. The same conclusion can be drawn for all three starting conditions, which indicates that the temperature rise of clutch 2 is generally higher than that of clutch 1.

Speed changes of the engine and clutch 2 after the pre-engagement phase are also analyzed. It is found that there is a proportional change between the engine and clutch 2 because the engine speed equals the speed of the wheel multiplied by the final drive ratio, and the speed of clutch 2 equals the speed of the wheel multiplied by the first gear ratio after the phase of pre-engagement. The speed difference therefore increases as the engine speed increases with the throttle opening, resulting in the temperature rise of clutch 2 being obviously higher than that of clutch 1. At the same time, the engine speed is higher for a fast start than for the other starting conditions, which makes the speed difference between the engine and the friction plate of clutch 2 obviously larger than that under the other starting conditions. Additionally, the grinding time of the two clutches is the same for the slow start and the normal start (i.e., 0.9 and 0.8 s, respectively), while there is a difference between the two clutches for the fast start (i.e., 0.8 s for clutch 1 and 0.95 s for clutch 2), with longer grinding producing more heat. In summary, the temperature rise of clutch 2 for a fast start is most serious.

6 Conclusion

According to three starting conditions (i.e., a slow start, normal start, and quick start), throttle acceleration simulations were carried out and the friction work and friction power of clutches 1 and 2 were calculated. The heat transfer coefficients of the pressure plate and friction plate were simulated for eight speeds. The temperature field of the two clutches was obtained using the calculated heat flux and convective heat transfer coefficient. The calculation results show that the temperature rise of clutch 2 is generally higher than that of clutch 1 for different starting conditions.
  1. (1)

    A constant-engine-speed control strategy was adopted for clutch 1 in the process of sliding, resulting in the speed difference between the engine and clutch 2 being higher generally than that between the engine and clutch 1. This led to the temperature rise of clutch 2 being higher generally.

     
  2. (2)

    The grinding temperature rise of clutch 2 can be further reduced if the control of the engine speed can be added to the shift process to minimize the speed difference. This has certain reference value for future research.

     

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Copyright information

© China Society of Automotive Engineers(China SAE) 2019

Authors and Affiliations

  1. 1.School of Automobile and Traffic EngineeringJiangsu UniversityZhenjiangChina
  2. 2.Shanghai Automobile Gear WorksShanghaiChina

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