TY - JOUR
AU - Zhang,, Jun
AB - Abstract The large longitudinal impact of heavy-haul trains is the main factor limiting their development, and the asynchronous nature of train-braking systems is the main cause of this longitudinal impact. In this paper, a segmented electro-pneumatic braking solution fully compatible with the existing freight-train braking system in China is proposed to improve the synchrony of train-braking systems. A simulation model for this braking system is developed based on air-flow theory, the 120 distribution valve and electronic control devices. The braking characteristics obtained from simulations are compared to those from the train-brake testing platform, and show high fidelity. On this basis, the effects of the new braking system on the braking capacity and longitudinal impact of a 20 000 t heavy-haul train are analysed by further simulation. The results show that during service brakes, the segmented electro-pneumatic braking system can increase the braking capacity by 4.2–24.7% and reduce the coupler force by 21.6–68.0%. Therefore, it can be seen that the segmented electro-pneumatic braking system is a new type of electro-pneumatic brake that meets the needs of the Chinese railway network. It solves the problem of the longitudinal impact of heavy-haul trains satisfactorily, and its compatibility with the existing braking system (resulting in a reduced modification workload) makes it possible to maintain normal operations on heavy-haul lines while trains undergo modification. 1. Introduction Because of its high efficiency and low cost, the heavy-haul train is widely used in railway freight transportation across the world, especially in the long-distance transportation of bulk goods such as ore and coal. However, as a concomitant of its unusual length and weight, the longitudinal impact of this type of train is significantly larger, leading to earlier damage to couplers and related parts and even coupler-breakage accidents, which threaten the safe operation of the train. Therefore, reducing longitudinal impact is one of the key research topics relating to heavy-haul trains. The main cause of longitudinal impact is the asynchronous nature of the pneumatic braking system. The brake wave in the pneumatic braking system depends on the propagation of air from the front to the back of the train. When the train brakes, the front vehicles brake first and the rear vehicles brake later, so the rear vehicles will cause longitudinal impact due to inertia. The air-wave propagation speed is about 320 m/s (approximately the speed of sound), so a braking system relying on pure air transmission is limited by the speed of air transmission, and it is impossible to solve the problem of longitudinal impact completely by increasing the braking-wave speed in the pneumatic braking system. In order to solve the problem of braking-system asynchrony, the United States and other countries have developed an electronically controlled pneumatic (ECP) braking system, which has been widely used in many countries and has achieved good results [1]. The ECP system controls the air flow from the auxiliary reservoir to the brake cylinder, which makes it incompatible with the existing Chinese train-braking system. If we adopt the ECP system for the Chinese railway, with its more than 750 000 vehicles, we will face the problems of high cost and a long modification period, the problem that modified and unmodified vehicles cannot be combined during the transition, and the possible need to change the management mode of freight trains. As a result, the ECP braking system has not been applied in China. Segmented electro-pneumatic braking is a new electronically controlled pneumatic braking method. Its basic approach is to divide the train into several virtual segments. Without changing the original pneumatic braking system, only one electronic control device is added to each segment. The electronic control device acts after receiving the remote control signal from the locomotive, and controls air exhaust during braking and the connection between the accelerated-release reservoir and the train pipe during release. In contrast to the ECP system, the device does not directly control the auxiliary reservoir to feed air to the brake cylinder, so the new system is fully compatible with the current braking system in use in China, and has good safety standards. For instance, if an electronic control device is lost or damaged, only the lost or damaged device is affected, which will affect the air-exhaust speed only of vehicles within the same segment; even in extreme cases, where all electronic control devices are out of service, the train will still retain pure pneumatic braking capacity. Therefore, this method offers an absolute guarantee of train safety. Segmented electro-pneumatic braking relies on compressed air and electrical signals to transmit braking signals together, so the braking capacity and synchrony of trains will also change greatly; at the same time, its braking-wave propagation law is completely different from that of the pure pneumatic braking system or the ECP system in the United States, as both the braking-system propagation characteristics and the brake-cylinder pressure-variation characteristics will greatly affect the longitudinal impact of the train. Thus, it is necessary to analyse the braking capacity of the new braking system and its effect on the longitudinal impact of trains. The simulation of the pneumatic train-braking system in the United States began in the 1970s, when a pneumatic braking-system model consisting of an ABD/ABDW valve and a 26C locomotive valve was initially developed [2]; the models of the pipeline system and the 26C locomotive valve were improved in 1986 [3], and the simulation system for locomotives and pneumatic braking was redeveloped in 2012 [4]. In 1985, Japan carried out a study on the reduction characteristics of the train-pipe model and models including branch pipes by developing models and using numerical calculation [5]. India also completed simulations of a vacuum braking system at the same time [6]; its model included the train pipe, the auxiliary reservoir and the brake cylinder. South Korea, Italy, Poland and a number of other countries have also carried out simulation research on braking systems [7–10]. China established a model involving the train pipe and branch pipe in the 1990s and analysed the impacts of various parameters on the reduction characteristics of the pipeline system [11]. Subsequently, a simulation model of a pneumatic braking system composed of GK, 104, F8, KZ1 and 120 distribution valves used in Chinese freight vehicles was established [12–17]. This study develops simulation models of train-braking and longitudinal train-dynamics systems, and, based on the characteristics of the new braking system obtained from a stationary testing platform, determines the difference in longitudinal impact via simulation. This is of great significance for further improving the design of the segmented electro-pneumatic braking system and guiding the study of longitudinal impact testing. 2. Working principles of the segmented electro-pneumatic braking system The pneumatic train-braking system is illustrated in Fig. 1. This braking system consists of two parts, namely the locomotive pneumatic brake unit and the vehicle pneumatic brake unit. The locomotive pneumatic brake unit includes the main reservoir, the compressor and the independent locomotive pneumatic braking system; the vehicle pneumatic brake unit includes the main train pipe, the branch pipe, connecting pipes between each reservoir and the distribution valve, the distribution valve, the auxiliary reservoir, the brake cylinder and the accelerated-release reservoir. Fig. 1. Open in new tabDownload slide Schematic diagram of the pneumatic train braking system Fig. 1. Open in new tabDownload slide Schematic diagram of the pneumatic train braking system The segmented electro-pneumatic braking system, as shown in Fig. 2, adds a set of electronic control devices to the original pneumatic braking system. During service and emergency brakes, the locomotive will issue wireless control instructions containing the value of reduction. The electronic control device will carry out air exhaust after receiving the wireless control braking signal, accelerate the pressure drop of the train pipe and close the exhaust port when the train-pipe pressure reduction reaches the value specified in the locomotive control command. During an emergency brake, the electronic control device still exhausts air at the service braking rate. At the same time, in order to minimize the impact of the number of vehicles in a segment on the exhaust speed of the exhaust device and to prevent excessively fast air exhaust in shorter segments and excessively slow air exhaust in longer segments, the electronic control device uses a secondary exhaust device, that is, a balancing reservoir is set up. After receiving the air-exhaust instruction, the electrically controlled air-exhaust device will first exhaust the compressed air in the balancing reservoir, and the relay valve will control the air exhaust of the train pipe according to the pressure difference between the balancing reservoir and the train pipe. When the locomotive sends out the release signal, the electronic control device will connect the train pipe through the upper chamber of the main valve with the balancing reservoir, so as to recharge the balancing reservoir and ensure normal air pressure in the balancing reservoir and the normal function of the relay valve for the next braking operation; as the balancing reservoir is recharged, the electronic control device connects the accelerated-release reservoir and the train pipe, thus realizing the accelerated-release effect. Fig. 2. Open in new tabDownload slide Schematic diagram of the segmented electro-pneumatic braking system Fig. 2. Open in new tabDownload slide Schematic diagram of the segmented electro-pneumatic braking system 3. Compressed air-flow equations in pipes Based on the characteristics of the pipeline of the braking system, it is assumed that the air flow in the braking system is one-dimensional, frictional and of unequal entropy flow. Based on the conservation of gas mass, momentum and energy, the equation system of the air-flow state can be obtained as follows: $$\begin{eqnarray*} \left\{ \begin{array}{@{}*{1}{l}@{}} {\displaystyle\frac{{\partial \rho }}{{\partial t}} + \rho \frac{{\partial u}}{{\partial x}} + u\frac{{\partial \rho }}{{\partial x}} + \frac{{\rho u}}{F}\frac{{dF}}{{dx}} = 0}\\ {\displaystyle\frac{{\partial u}}{{\partial t}} + u\frac{{\partial u}}{{\partial x}} + \frac{1}{\rho }\frac{{\partial p}}{{\partial x}} + \frac{{4f}}{D}\frac{{{u^2}}}{2}\frac{u}{{\left| u \right|}} = 0}\\ {\displaystyle\frac{{\partial p}}{{\partial t}} + u\frac{{\partial p}}{{\partial x}} - {a^2}\frac{{\partial \rho }}{{\partial t}} - {a^2}u\frac{{\partial \rho }}{{\partial x}}}\\ \,\,\,\,\,\,\quad -\, (k - 1)\rho \left(q + u\displaystyle\frac{{4f}}{D}\frac{{{u^2}}}{2}\frac{u}{{\left| u \right|}}\right) = 0 \end{array} \right. \end{eqnarray*}$$(1) In these equations, ρ,u and p are the gas density, velocity and pressure, respectively; x, F, D and f are the distance, tube area, tube diameter and friction coefficient of the inner wall, respectively; t is the time, a is the sound velocity; and k and q are the gas-specific heat ratio and the heat exchange quantity per unit time, respectively. The above partial differential equations have no analytical solution and need to be solved by the numerical method. In this paper, the characteristic line method is used [11]. The various gas physical quantities are converted into Riemann variables and entropy expressions, and three characteristic line directions and corresponding incremental expressions are obtained after dimensionless transformation: Along the |$\lambda $| characteristic line: $$\begin{eqnarray*} d\lambda &=& \frac{A}{{{A_A}}}d{A_A} - \frac{{k - 1}}{2}\frac{{AU}}{F}\frac{{dF}}{{dX}}dZ + \frac{{{{(k - 1)}^2}}}{2}\frac{{q{x_{ref}}}}{{{a^3}_{ref}}}\frac{1}{A}dZ\nonumber\\ &&-\, \frac{{k - 1}}{2}\frac{{2f{x_{ref}}}}{D}{U^2}\frac{U}{{\left| U \right|}}\left[1 - (k - 1)\frac{U}{A}\right]dZ \end{eqnarray*}$$(2) Along the |$\beta $| characteristic line: $$\begin{eqnarray*} d\beta &=& \frac{A}{{{A_A}}}d{A_A} - \frac{{k - 1}}{2}\frac{{AU}}{F}\frac{{dF}}{{dX}}dZ + \frac{{{{(k - 1)}^2}}}{2}\frac{{q{x_{ref}}}}{{{a^3}_{ref}}}\frac{1}{A}dZ\nonumber\\ &&+\, \frac{{k - 1}}{2}\frac{{2f{x_{ref}}}}{D}{U^2}\frac{U}{{\left| U \right|}}\left[1 + (k - 1)\frac{U}{A}\right]dZ \end{eqnarray*}$$(3) The characteristic line along the entropy trace: $$\begin{eqnarray*} d{A_A} = \frac{{k - 1}}{2}\frac{{{A_A}}}{{{A^2}}}\left[\frac{{q{x_{ref}}}}{{{a_{ref}}}} + \frac{{2f{x_{ref}}}}{D}{\left| U \right|^3}\right]dZ \end{eqnarray*}$$(4) In these equations, λ and β are Riemann variables, AA is entropy, X is the dimensionless distance, Z is the dimensionless time, U is the dimensionless velocity, A is the dimensionless sound velocity, k is the specific heat ratio, f is the tube-wall friction, F is the dimensionless area, xref is the reference length and aref is the reference sound velocity. The above equations can only solve the gas state in the pipeline; for the pipeline boundary areas, additional boundary equations are required. The boundary equations used in this system include partial open-end, multi-pipe branch boundary, pipe-cylinder connection boundary and so on [16]. 4. Physical model of the distribution valve and electronic control device In a train with a segmented electro-pneumatic braking system, the braking system of some vehicles is equipped with an electronic control device. The braking system of vehicles without the electronic control device can be represented with the existing vehicle braking system model [17], while the braking system of vehicles with the electronic control device requires the electronic control-device model to be added to the original model, as shown in Fig. 3. Fig. 3. Open in new tabDownload slide Model of a vehicle braking system with electronic control Fig. 3. Open in new tabDownload slide Model of a vehicle braking system with electronic control The vehicle braking system model consists of two main pipes (the main pipe is divided into two sections at the connection between the main pipe and the branch pipe) and one branch pipe; three inter-cylinder connecting pipes, which connect the lower chamber of the main valve to the brake cylinder, auxiliary reservoir and accelerated-release reservoir; seven chambers, which are the upper and lower chambers of the 120 distribution valve, the brake cylinder, the auxiliary reservoir, the accelerated-release reservoir, the emergency room and the balancing reservoir of the electronic device. With the exception of ϕ10-ϕ12, which are the orifices in the electronic control device, the diameter of the other nine orifices are all related to the state of the 120 distribution valve, that is, related to the position of the moving parts in the distribution valve. In the process of braking and release, the sizes of these orifices are constantly changing, and their opening conditions and sizes have been described in detail in an earlier study [17]; this paper will focus on the working principles of the electronic control device. The electronic control signal of the locomotive directly controls the area of ϕ10, the exhaust orifice of the balancing reservoir in the electronic control device, and then determines the working state of the relay valve according to the pressure difference of the balancing reservoir and the upper chamber of the distribution valve (connected with the train pipe). When the pressure difference reaches the working condition, ϕ11 is opened, thus the compressed air in the train pipe is exhausted into the atmosphere. The exhaust-termination pressure of the balancing reservoir is controlled by the locomotive control signal, which is equal to the reduction applied by the driver. When the pressure reduction of the balancing reservoir reaches the amount set by the locomotive instruction, the balancing reservoir exhaust port ϕ10 is closed. When the exhaust orifice ϕ11 of the train pipe is opened, the pressure difference between the balancing reservoir and the train pipe will decrease continuously as the pressure of the train pipe decreases; when the two are balanced, the relay valve closes the ϕ11 orifice and the remote control reduction (to distinguish it from the local reduction of the train pipe and the reduction of the train pipe caused by the locomotive exhaust port) of the train pipe stops. When the electronic control device receives the locomotive release remote-control instruction, it will open the ϕ12 orifice, which is connected with the balancing reservoir through a virtual pipeline (the dashed line in Fig. 3). At this time, the air in the train pipe will flow into the balancing reservoir through the upper chamber of the distribution valve to recharge the balancing reservoir. Finally, when the pressure of the balancing reservoir and the train pipe is balanced, the ϕ12 orifice of the balancing reservoir is closed, and the balancing reservoir stops recharging. After receiving the release instruction, the electronic control device will connect the accelerated-release reservoir and the upper chamber of the main valve, so the air in the accelerated-release reservoir can flow into the upper chamber of the main valve and enter the train pipe, in order to realize accelerated release. The electronic control device of the segmented electro-pneumatic braking system involves the movement of mechanical parts after receiving electronic signals to realize the air-exhaust function. In theory, the transmission time of the electronic signals is negligible, but considering the lag of the mechanical system, the possible attenuation during electronic signal transmission and the possibility that the electronic control device may start to work only when the second or later signal is received, this simulation system assumes that there are three situations for the initial action of the electronic control device (when the balancing reservoir starts to exhaust): first, synchronous air exhaust without time difference for all electronic control devices, that is, all electronic control devices work synchronously with the locomotive; second, each electronic control device lags in accordance with the law of normal random distribution, and the time range of random variation is variable; and third, the devices act sequentially at a determined wave speed, that is, in the form of a uniform propagation towards the rear end. 5. Results of the stationary brake test The electronic control device of the vehicle braking system is developed by Meishan Brake Science & Technology in accordance with the control principles outlined above, and braking-performance testing is carried out on 150 brake-test platforms (as shown in Fig. 4) at the company's factory. The braking performance of a 108-vehicle marshalled train was tested for this study. An electronic control device was installed for each vehicle, with a total of 10 test sections located in vehicles 1, 10, 20, 30, 40, 50, 58, 68, 78, 88, 98 and 108. The pressure of the train pipe, the pressure of the auxiliary reservoir and the pressure of the brake cylinder were recorded under conditions of braking, release and staged braking at different levels of pressure reduction. Because the brake cylinder pressure has a direct impact on the longitudinal impact of the train, this paper only gives the results of several brake-cylinder pressure tests and the corresponding simulation results. Brake-cylinder testing and simulation results for a reduction of 50 kPa are shown in Figs 5 and 6. Fig. 4. Open in new tabDownload slide Brake-test platform at Meishan Brake Science & Technology Fig. 4. Open in new tabDownload slide Brake-test platform at Meishan Brake Science & Technology Fig. 5. Open in new tabDownload slide Brake-cylinder testing results for a 50 kPa reduction Fig. 5. Open in new tabDownload slide Brake-cylinder testing results for a 50 kPa reduction Fig. 6. Open in new tabDownload slide Brake-cylinder simulation results for a 50 kPa reduction Fig. 6. Open in new tabDownload slide Brake-cylinder simulation results for a 50 kPa reduction From Fig. 5, it can be seen that when braking at a 50 kPa reduction, the braking start-time difference between the first and last vehicles (calculated from the beginning of the brake-cylinder inflation) was only 1.2 seconds, the braking synchronization was very good, and the operation was basically transmitted from the front to the back in each vehicle. At the initial inflation, there was a pause in the rise of the brake-cylinder pressure at about 40–50 kPa. This corresponds to the brake-cylinder piston protruding action, after which the brake-cylinder pressure slowly rose. The rising rate of the brake-cylinder pressure increase of each vehicle was slightly different, and the brake-cylinder pressure of some vehicles appears to have been in a jagged shape during the rising process, which signifies that the corresponding distribution valve was in the process of braking and pressure-holding transformation, that is, the opening and closing of the air-feeding orifice of the brake cylinder. The average maximum rising pressure of the brake cylinder was 120 kPa, and the maximum pressure varied from vehicle to vehicle, with a maximum value of 132 kPa and a minimum value of 108 kPa. The release test began at 60 seconds. During the release process, the brake cylinders exhausted at approximately the same rate, and the difference in the release-start times of the first and last vehicles was about 2 seconds. As the brake-cylinder pressure dropped to about 13 kPa, the brake-cylinder piston retracted, and then the exhaust speed slowed down significantly. The simulation results in Fig. 6 show an obvious step on the curve at 40 kPa, suggesting that the protruding pressure of the brake-cylinder piston was 40 kPa. The results are similar to the testing results in terms of the time difference between the first and last vehicles, the rate of pressure increase in the brake cylinder, the jagged shape during the rise and the final balanced pressure. Testing and simulation results for the brake-cylinder pressure change for a 170 kPa reduction are shown in Figs 7 and 8. The testing values of the braking start-time difference (calculated from the beginning of the brake-cylinder inflation) and the release-time difference between the first and last vehicles were 2.5 seconds and 2.5 seconds, respectively; the corresponding simulation results were 2.8 seconds and 2.8 seconds. The shape of the pressure-rise curve of the brake cylinder was basically the same in both testing and simulation, and the balanced pressure of the brake cylinder in testing and simulation was 460 kPa and 463 kPa, respectively. Thus, the simulation and testing results have a good degree of agreement. Therefore, based on the agreement of the above simulation and testing results, it can be asserted that the braking-system simulation model has high fidelity, and meets the basic conditions for longitudinal train-impact simulation. Fig. 7. Open in new tabDownload slide Brake-cylinder testing results for a 170 kPa reduction Fig. 7. Open in new tabDownload slide Brake-cylinder testing results for a 170 kPa reduction Fig. 8. Open in new tabDownload slide Brake-cylinder simulation results for a 170 kPa reduction Fig. 8. Open in new tabDownload slide Brake-cylinder simulation results for a 170 kPa reduction 6. Prediction of braking and longitudinal dynamics of trains On the basis of the accurate simulation of the characteristics of brake-cylinder pressure changes in the train-braking system, the differences between the electro-pneumatic braking system and the pure pneumatic braking system in terms of braking capacity and longitudinal impact can be predicted by the simulation system. The train air-braking and longitudinal dynamics simulation system (TABLDSS) used in this study has a high level of accuracy in train-operation process and coupler-force prediction [18]. The longitudinal force was calculated by taking the 20 000 t train—which is widely used on heavy-haul lines and shows a relatively prominent problem of longitudinal impact—as an example. The train marshalling was 1 locomotive (HXD1) + 108 vehicles (C80) + 1 locomotive (HXD1) + 108 vehicles (C80) + controllable EOT. During the simulation, the initial speed of the train was 70 km/h; all the vehicles were equipped with the electronic braking-control device, and both the slave locomotive and the controllable EOT lagged behind the main control locomotive by 2 seconds. Service braking conditions for reductions of 50 kPa, 100 kPa and 170 kPa were calculated. The speed change and running distance of the 20 000 t train for a service-brake reduction of 100 kPa are shown in Fig. 9. In order to illustrate the difference in braking capacity compared with the pure pneumatic braking system, the corresponding curve under pure pneumatic-braking mode is also shown in Fig. 9. From the two speed curves in Fig. 9, it can be seen that the electro-pneumatic braking speed dropped noticeably faster, dropping to zero at 74 seconds. At this time, the train running distance was 903 m. In comparison, the speed of the pneumatic brake dropped to zero at 94 seconds and the running distance was 1199 m. Therefore, the stopping distance of the electro-pneumatic brake was shorter by 24.7% compared with that of the pure pneumatic brake. The braking-distance changes under other pressure-reduction conditions are shown in Table 1. Fig. 9. Open in new tabDownload slide Braking distance and speed of the pneumatic and electro-pneumatic braking systems for a 100 kPa reduction Fig. 9. Open in new tabDownload slide Braking distance and speed of the pneumatic and electro-pneumatic braking systems for a 100 kPa reduction Table 1. Comparison of pneumatic and electro-pneumatic braking distances Pressure reduction (kPa) . Pneumatic braking distance (m) . Electro-pneumatic braking distance (m) . Reduction (%) . 50 2737 2623 4.2 100 1199 903 24.7 170 889 687 22.7 Pressure reduction (kPa) . Pneumatic braking distance (m) . Electro-pneumatic braking distance (m) . Reduction (%) . 50 2737 2623 4.2 100 1199 903 24.7 170 889 687 22.7 Open in new tab Table 1. Comparison of pneumatic and electro-pneumatic braking distances Pressure reduction (kPa) . Pneumatic braking distance (m) . Electro-pneumatic braking distance (m) . Reduction (%) . 50 2737 2623 4.2 100 1199 903 24.7 170 889 687 22.7 Pressure reduction (kPa) . Pneumatic braking distance (m) . Electro-pneumatic braking distance (m) . Reduction (%) . 50 2737 2623 4.2 100 1199 903 24.7 170 889 687 22.7 Open in new tab It can be seen from Table 1 that electro-pneumatic braking was able to shorten the braking distance. The braking distance was shortened by 4.2% when at 50 kPa of pressure reduction; as the pressure reduction increased, the braking distance was shortened more significantly, and when the pressure reduction was large, the braking distance was shortened by more than 20%. It was also found that the consistency of the shape of the braking-pressure curve of the train vehicles was the main reason for the variation in braking capacity. When the pneumatic brake was applied, the consistency of the shape of the braking-pressure curve of each vehicle's brake cylinder was poor, while the electro-pneumatic brake produced the opposite result. Take the 170 kPa reduction as an example: the brake cylinder of the front vehicle reached the maximum pressure within 20 seconds after application of the pneumatic brake, while the rear vehicle reached the maximum pressure only about 100 seconds after application of the brake, which is quite an obvious difference. With the electro-pneumatic brake, the shape of the pressure curve of each vehicle's brake cylinder was almost identical. Therefore, in the case of the large pressure reduction, the consistency of the pressure-curve shape of the electro-pneumatic brake cylinders was much better than that of the pneumatic brake cylinders, which was the main reason for the difference in braking capacity. Fig. 10 shows the distribution of the maximum coupler-compression force of each vehicle along the length of the train during braking with a reduction of 170 kPa. The maximum coupler-compression force here refers to the maximum coupler-compression force that each vehicle bore during braking. The coupler-force distribution of the two braking modes—pneumatic and electro-pneumatic—is shown in the figure. During service braking, the maximum coupler force of the pneumatic brake appeared in the middle vehicle, and throughout the whole train, the coupler compression force was basically distributed in a 'fish belly' shape. The maximum coupler force (−1919 kN) occurred in vehicle 125. When using the segmented electro-pneumatic braking system, the maximum coupler-compression force was −615 kN and occurred in vehicle 109. Compared with the traditional pneumatic braking system, the maximum coupler force of the train decreased by 68%. It can be seen that the level of the coupler force can be significantly reduced using the segmented electro-pneumatic braking system. Compared with the pneumatic braking system, for the segmented electro-pneumatic braking system, the consistency of the action-time difference between the front and rear vehicles and the shape of the pressure curve of the brake cylinder of each vehicle were markedly improved, so the coupler force was significantly reduced. The sudden change in the coupler force at around vehicle 110 was due to the fact that that car was the slave locomotive, and its ratio of air-brake force and weight differed from those of the adjacent vehicles. Fig. 10. Open in new tabDownload slide Comparison of maximum coupler forces in service braking for a 170 kPa reduction Fig. 10. Open in new tabDownload slide Comparison of maximum coupler forces in service braking for a 170 kPa reduction The changes in the coupler force of the 20 000 t train under various pressure-reduction conditions are shown in Table 2. Table 2. Comparison of coupler forces with pneumatic and electro-pneumatic breaking Pressure reduction (kPa) . Maximum coupler force with pneumatic braking (kN) . Maximum coupler force with electro-pneumatic braking (kN) . Reduction (%) . 50 −320 −251 21.6 100 −1354 −544 59.8 170 −1919 −615 68.0 Pressure reduction (kPa) . Maximum coupler force with pneumatic braking (kN) . Maximum coupler force with electro-pneumatic braking (kN) . Reduction (%) . 50 −320 −251 21.6 100 −1354 −544 59.8 170 −1919 −615 68.0 Open in new tab Table 2. Comparison of coupler forces with pneumatic and electro-pneumatic breaking Pressure reduction (kPa) . Maximum coupler force with pneumatic braking (kN) . Maximum coupler force with electro-pneumatic braking (kN) . Reduction (%) . 50 −320 −251 21.6 100 −1354 −544 59.8 170 −1919 −615 68.0 Pressure reduction (kPa) . Maximum coupler force with pneumatic braking (kN) . Maximum coupler force with electro-pneumatic braking (kN) . Reduction (%) . 50 −320 −251 21.6 100 −1354 −544 59.8 170 −1919 −615 68.0 Open in new tab During braking of the 20 000 t train with the segmented electro-pneumatic braking system, therefore, the compressed coupler force was greatly improved, and as the pressure reduction increased, the reduction in the coupler force was more and more significant. 7. Conclusions This paper proposed a segmented electro-pneumatic braking system that is suitable for China's national conditions. The system is a new braking mode that takes the electronic control of exhaust as the control target and is completely different from the existing ECP direct-control system. First, the simulation model of the segmented electro-pneumatic braking system was developed based on air-flow theory and the controlling principle of distribution valves and electronic control devices, and the braking capacity and longitudinal impact of a 20 000 t train were predicted. The results showed that segmented electro-pneumatic braking clearly increases the speed of braking-wave propagation, and at the same time accelerates the increase in brake-cylinder pressure in the rear vehicle. The shape of the pressure curve of each vehicle's brake cylinder in the train is more consistent, which improves the braking synchrony of the train. The braking capacity of the train is enhanced when the segmented electro-pneumatic braking mode is adopted. In the case of service braking, the braking capacity is increased by about 4% when the pressure reduction is small, and by about 20% when the pressure reduction is large. Because the increase of synchrony offered by the segmented electro-pneumatic braking system has obvious advantages for reducing the longitudinal impact of the train; when the service brake in a 20 000 t combined train is used, the coupler force is reduced by a minimum of 21.6%, and the maximum coupler force decreases by 68% as the pressure reduction increases. As a result, the segmented electro-pneumatic braking system is a good way to solve the problem of longitudinal impact in heavy-haul trains. Conflict of interest statement None declared. References 1. Liu Y , Zhang G, Zhang S, et al. . Feasibility to develop ECP braking technology for railway in China . Rolling Stock . 2014 ; 52 : 29 – 32 . Google Scholar OpenURL Placeholder Text WorldCat 2. Gauthier RG . An analysis and simulation of a pneumatic control valve system . Master's Thesis . University of New Hampshire 1977 . Google Scholar Google Preview OpenURL Placeholder Text WorldCat COPAC 3. Abdol-Hamid KS . Analysis and simulation of the pneumatic braking system of freight trains . Ph.D. Thesis . University of New Hampshire 1986 . Google Scholar Google Preview OpenURL Placeholder Text WorldCat COPAC 4. Specchia S , Afshari A, Shabana AA, et al. . A train air brake force model: locomotive automatic brake valve and brake pipe flow formulations . P I Mech Eng F J Rai . 2012 ; 227 : 19 – 37 . Google Scholar OpenURL Placeholder Text WorldCat 5. Tanaka H , Hasegawa I. Characteristics of pressure reduction in air brake pipe of rolling stock . Railway Tech Res Inst Q Rep . 1986 ; 27 : 127 – 9 . Google Scholar OpenURL Placeholder Text WorldCat 6. Bharath S , Nakra BC, Gupta KN. Mathematical model of a railway pneumatic brake system with varying cylinder capacity effects . J Dyn Syst-T ASME . 1990 ; 112 : 456 – 62 . Google Scholar Crossref Search ADS WorldCat 7. Nam SW , Kim H. A study on the improvement of release application characteristics of pneumatic brakes for freight train . KSME Int J . 2002 ; 16 : 776 – 84 . Google Scholar Crossref Search ADS WorldCat 8. Pugi L , Malvezzi M, Allotta B, et al. . A parametric library for the simulation of a Union Internationale des Chemins de Fer (UIC) pneumatic braking system . P I Mech Eng F J Rai . 2004 ; 218 : 117 – 32 . Google Scholar OpenURL Placeholder Text WorldCat 9. Cantone L , Crescentini E, Verzicco R, et al. . A numerical model for the analysis of unsteady train braking and releasing manoeuvres . P I Mech Eng F J Rai . 2009 ; 223 : 305 – 17 . Google Scholar OpenURL Placeholder Text WorldCat 10. Piechowiak T . Pneumatic train brake simulation method . Veh Syst Dyn . 2009 ; 47 : 1473 – 92 . Google Scholar Crossref Search ADS WorldCat 11. Wei W , Zhang S, Liu Q. A study on characteristic of pressure reduction of air brake system in a long train . Dalian Jiaotong Univ . 1992 ; 13 : 43 – 49 . Google Scholar OpenURL Placeholder Text WorldCat 12. Wei W , Zhang K. A mathematical model of air brake system of train . J Southwest Jiaotong Univ . 1994 ; 29 : 283 – 91 . Google Scholar OpenURL Placeholder Text WorldCat 13. Xu Y . Brake system performance predict for express freight train with 104 distribution valve . Railway Locom Car . 2007 ; 7 : 21 – 26 . Google Scholar OpenURL Placeholder Text WorldCat 14. Wei W , Liang D. Simulation and analysis of the brake system of express freight train with F8 distribution valve . Railway Locom Car . 2007 ; 27 : 142 – 5 . Google Scholar OpenURL Placeholder Text WorldCat 15. Wei W , Tao L, Jun Z. The simulation model of KZ1 control valve and the simulation study on train braking performance . China Railway Sci . 2010 ; 31 : 105 – 9 . Google Scholar OpenURL Placeholder Text WorldCat 16. Wei W , Hu Y, Wu Q, et al. . An air brake model for longitudinal train dynamics studies . Veh Syst Dyn . 2017 ; 55 : 517 – 33 . Google Scholar Crossref Search ADS WorldCat 17. Wei W , Ye L. Simulation of a freight train brake system with 120 valves . P I Mech Eng F J Rai . 2009 ; 223 : 85 – 92 . Google Scholar OpenURL Placeholder Text WorldCat 18. Qing W , Spiryagin M, Colin C, et al. . International benchmarking of longitudinal train dynamicssimulators: results . Veh Syst Dyn . 2018 ; 56 : 343 – 65 . Google Scholar Crossref Search ADS WorldCat © The Author(s) 2020. Published by Oxford University Press on behalf of Central South University Press. This is an Open Access article distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/by/4.0/), which permits unrestricted reuse, distribution, and reproduction in any medium, provided the original work is properly cited.
TI - Study on a segmented electro-pneumatic braking system for heavy-haul trains
JF - Transportation Safety and Open Environment
DO - 10.1093/tse/tdaa015
DA - 2020-10-27
UR - https://www.deepdyve.com/lp/oxford-university-press/study-on-a-segmented-electro-pneumatic-braking-system-for-heavy-haul-VKFX54erro
SP - 216
EP - 225
VL - 2
IS - 3
DP - DeepDyve
ER -