PAPERS
Experiments on Damage to Track Components due to Repeated Passage of Vehicles on Rail Gaps
2025 Volume 66 Issue 1 Pages 23-30
Details
2025 Volume 66 Issue 1 Pages 23-30
In railways where wireless train control systems are employed, track circuits may be removed, making it difficult to detect a broken rail in such cases. In railways without track circuits, it is assumed that vehicles repeatedly pass over rail gaps until the broken rail is detected by rail inspection or other means. Therefore, in order to evaluate damage of track components due to repeated vehicle passage, we carried out falling weight tests in a laboratory, in which impact loads equivalent to passing vehicles were applied to the broken rail to simulate damage. This test clarified the plastic deformation of the rails and functional deterioration of the rail fastenings system in response to the impact loads.
Track circuits are the main method for detecting the position of a vehicle on a railway. This method detects that a vehicle is on a particular section of track by using the fact that the current flow on the rails changes depending on whether a wheel set is on the rails or not. In addition, the track circuit also has the function of immediately detecting a rail break, because the way the current flows changes when the rail has been broken, creating a gap. In this study, a rail gap is defined as a point where the rail has been broken creating a gap. On the other hand, in recent years, wireless train position detection methods (hereinafter referred to as “wireless train control systems”) have been introduced overseas [1], and some Japanese railway operators are also introducing such systems [2]. Wireless train control systems detect the position of a train using radio equipment mounted on the train. This system allows greater flexibility in operation management compared to the section-by-section detection method using track circuits. The wireless train control system also has the advantage of reducing maintenance costs by streamlining the above-ground equipment used in current track circuits. However, if the track circuit is removed due to the introduction of a wireless control system, it will no longer possible to detect broken rails using the current method. Therefore, alternative systems for detecting broken rails, such as on-board detection methods, are being researched [3]. However, unlike track circuits that can constantly monitor rail defects, there is a possibility that vehicles may repeatedly run over a rail gap at normal operating speeds before the broken rail is detected. Previous studies have investigated the running safety of a single vehicle passing over a rail gap [4]. However, assuming a situation where vehicles repeatedly pass over the rail gap, the track components around the rail gap are repeatedly subjected to impact loads, which may reduce running safety due to the damage to these components.
Previous studies have examined the damage to track components due to impact loads caused by vehicles repeatedly passing over rail gaps. These studied rail damage caused by approximately 600 repeated loadings [5], but rail damage caused by further repeated loading and damage to other track components has not been clarified. In this study, to clarify the damage to the components such as rails and rail fastening systems caused by impact loads during repeated vehicle passage over rail gaps, we developed and carried out tests with a falling weight test device capable of simulating the impact loads of a vehicle passing over a rail gap.
In this study, we assume a situation in which vehicles repeatedly pass over the rail gap for a certain period of time between the time the rail is broken and the time the rail gap is detected. As a specific example of this situation, assuming an urban/suburban railway line operating 6 trains per hour each with 10 vehicles per train (4 axles per vehicle) for 18 hours per day, 4,320 axles will pass over the rail gap in one day. Furthermore, assuming that the inspection cycle for reliable detection of rail gaps runs in a seven-day cycle carried out by foot patrols along the track, it is estimated that 30,240 axles will pass over the track during this period. It is therefore necessary to check the occurrence and progression of damage to track components under such conditions. Methods to evaluate the damage to track components at the rail gap include test runs with real vehicles, repeated calculations by numerical analysis and element tests simulating impact loads. However, assuming the number of repetitions mentioned above, running test with a real vehicle is not realistic from the viewpoint of safety, and numerical analysis is not realistic from the viewpoint of computational capacity. Therefore, element tests that apply repeated impact loads during passage through the rail gap are currently considered the most appropriate for evaluation. In addition, it is considered appropriate to use a falling weight test device capable of repeated impact loading for the element test. On the other hand, it is expected that a large test device will be required to apply an impact load simulating passage of a vehicle over the gap. In a large-scale falling weight test device, the weight is often hoisted or dropped using chains or wires to lift the weight and then release it in free fall. It is expected that achieving a load cycle count in the tens of thousands will require a long period of time. In this study, we developed a new falling weight test device that can automatically apply repeated and controlled impact loads in the order of tens of thousands to reproduce passage over the rail gap. By using a cam mechanism to lift the weight, we enabled automatic repeated loading, and by using an inverter motor, we can set shorter loading intervals.
2.2 Calculation of required capacity of test device by running analysis on rail gapIn order to investigate the required loads for the test device, the impact loads generated by vehicles passing over the rail gap were calculated using a dynamic FEM analysis method [6]. The analysis model is shown in Fig. 1. The track is shown in Figs. 1(a) and 1(d). The rails are modeled with beams and rigid elements, the sleepers with beam elements and the roadbed with spring and damper elements. As shown in Fig. 1(c), the vehicle model is a three-dimensional model in which the car body, bogie, and wheelset are modeled as rigid bodies, and coupled with spring and damper elements. Table 1 shows the main specifications of the track model and the vehicle model. Using this analysis model, the impact load at the time of passing over the rail gap was calculated, and this load was used as the required capacity of the falling weight test device. It is assumed that the damage to the components caused by the impact load is influenced by the value of the impact load and its duration, i.e., the impulse. Therefore, each value was calculated accordingly.
| (a) Track | |
| Item | Specifications |
| Track gauge | 1067 mm |
| Rail type | JIS 50 kgN Rail |
| Rail fastenings vertical springs | 110 kN/mm |
| Roadbed vertical springs | 3.3×105 N/mm |
| (b) Vehicle | |
| Item | Specifications |
| Body mass | 28.3 tonne |
| Mass between primary spring and secondary spring | 2.2 tonne/bogie |
| Unsuspended mass | 1.52 tonne/wheelset |
| Wheelbase | 2.1 m |
| Distance between two bogies | 14.15 m |
| Shape of wheel surface | Modified arc wheel profile (JIS) |
Table 2 shows the track conditions assumed in this study. In this study, we assumed a rail gap on a well-maintained ballastless track on a conventional line. We set the condition where the rail gap is located at the end of the sleeper on the receiving side, as this is where significant wheel loads are likely to occur. The “leaving side” refers to the section of rail leading up to the gap. The “receiving side” refers to where the wheel runs onto the next section of rail after the gap, as illustrated in Fig. 2, where the direction of running is left to right. The reason for setting these conditions is that when calculating the loading conditions for tests No. 1 to No. 3 of the falling weight tests described later in Chapter 3, the impact load was greatest for the condition (test No. 1) where the end of the rail gap was aligned with the edge of the sleeper on the receiving side in a straight line. Specifically, when the rail gap ends at the edge of the sleeper on the receiving side, as shown in Fig. 3, a significant height difference occurs between the leaving side rail and the receiving side rail as the wheel passes over the rail gap. This level difference is expected to cause a significant impact on the receiving rail. For the rails, of the widely used 50-60 kgN rails, we focused on the less rigid 50 kgN rail. There are many types of rail fastening systems, but we chose to use a bar-shaped spring clip type rail fastening system with a high rail fastening force. This is because, when focusing on rail damage, assuming that the spring constant of the elastic rail pad laid directly under the rail is the same, the greater the rail fastening force, the smaller the rail displacement under train loads. This is expected to reduce the dispersion effect of impact loads, making plastic deformation of the rail head more likely. The fastening interval was set at 750 mm, the largest interval among the track types used on typical conventional suburban lines.
| Item | Specifications |
| Track line | Straight (Speed 100 km/h) |
| Track structure | Ballastless track |
| Fastening interval | 750 mm |
| Rail pad vertical springs | 110 MN/m |
| Rail gap | 70 mm |
| Position of rail gap | End of the receiving sleeper |
Figure 4 shows the results of the running analysis under the conditions described above. The analysis showed that the maximum impact load acting on the rail when the wheel passed over the rail gap was 170 kN, the duration of action was 8 msec, and the impulse was 537.4 N·s. The impact load occurred only on the receiving rail. Based on these results, the components to be evaluated for damage in the falling weight test were the receiving track components. However, if the weight is dropped on to the rail end while only the track components on the receiving side are simulated, the behavior of the weight during the drop becomes unstable. This situation raises concerns about variations in the loading position and increased stress on the test device. Therefore, as shown in Fig. 5, the components on the receiving side were arranged so that they were paired around the rail gap, and the loading behavior was stabilized by dropping the weight on both rails at the same time. It was therefore decided that the falling weight test device should be capable of applying an impact load of 340 kN, which is twice 170 kN. Based on these conditions, the test device was manufactured.
Based on the studies carried out up to the previous section, the appearance of the manufactured test device is shown in Fig. 6 and its specifications are given in Table 3. In this test device, a cam system is used to lift the weight, and an inverter motor is used to adjust the loading cycle of the weight, so that the minimum loading cycle for stable loading is 2.5 seconds per cycle. In addition, to stabilize the unstable behavior of the falling weight test onto the rail gap, the weight is guided by two guide bars. Performance verification tests were carried out on a specimen simulating the rail gap using the test device. As a result, it was confirmed that the test device can be adjusted up to 346 kN, 9 msec, and 1,557 N·s of impulse, compared to the target load of 340 kN, action time of 8 msec, and 1,360 N·s of impulse. It was also confirmed that the impact load was stable in the range of +1 to +3% when continuously operated at the minimum loading period.
| Item | Specifications |
| Driving source | Inverter motor |
| Driving system | Cam mechanism |
| Operation system | Manual/Automatic |
| Manipulation system | Hand-held controller/ Control panel |
| Falling mass | 640~1,240 kg |
| Falling height | 25~80 mm |
| Falling cycle | 2.5~12.5 s/times |
Using the manufactured falling weight test device, we carried out damage verification tests on the track components, simulating the repeated passage of vehicles over the rail gap. The test conditions are given in Table 4. In this test, the rail and the rail fastening system were evaluated and three numbered tests were carried out for straight and curved running conditions. In each numbered test, the rail gap was positioned to end at the edge of the sleeper on the receiving side or begin at the edge of the sleeper on the leaving side, as shown in Fig. 7, in order to increase the load on the rail and the rail fastening system, respectively. In addition, the loading conditions for each numbered test were calculated for the impact load, duration of action, and its impulse using the rail gap running analysis method described in Section 2.2. Because of the characteristics of the test, it was difficult to match the impact load and duration of action exactly to the target values, so the load was adjusted so that the impulse was not less than the target value. The target number of load cycles was set at 100,000 cycles in order to check for component damage from further repeated impacts in addition to the repeated passage conditions described in Section 2.1. An example of specimen installation is shown in Fig. 8. In this test, the receiving track components were installed in pairs each side of the rail gap in order to stabilize the weight behavior as described in Fig. 5. For evaluation, measurements were performed on the rail and rail fastening system on the north side of the specimen, which was arranged symmetrically in a north-south direction around the rail gap, as shown in Fig. 8. The detailed conditions of each test number are described below.
| Item | Specifications | |
| Rail type | JIS 50 kgN Rail | |
| Rail fastening system | Bar-shaped spring clip | |
| Rail pad vertical springs | 110 MN/m | |
| Fastening interval | 750 mm | |
| Test number | 1 | Straight·test to confirm rail damage (rail gap begins at edge of receiving sleeper) |
| 2 | Straight·test to confirm rail fastening system damage (rail gap is at the end of the leaving sleeper) | |
| 3 | Curve·test to confirm rail fastening system damage (rail gap begins at edge of receiving sleeper) | |
| Test load (per rail), duration of action, impulse | Test number 1: 170 kN, 8 msec, 680 N·s | |
| Test number 2: 100 kN, 19 msec, 950 N·s | ||
| Test number 3: 98 kN, 23 msec, 1,127 N·s | ||
| Loading times | 100,000 times | |
Test No. 1 was a test to confirm rail damage, under the condition that the end of the rail gap is the edge of the receiving sleeper in a straight line. This is because when the end of the rail gap is aligned with the edge of the receiving sleeper (Fig. 7(a)), there is a step difference between the leaving and receiving rails, which is considered to place a large load on the receiving rail. The assumed running speed was 100 km/h. The target test load, duration of action, and impulse were 170 kN, 8 msec, and 680 N·s.
3.1.2 Test No. 2 (Straight line test to confirm rail fastening system damage)Test No. 2 was a test to confirm damage to the rail fastening system under the condition that the rail gap begins from the edge of the sleeper on the leaving side in a straight line. This is because, when the rail gap begins from the edge of the leaving side sleeper (Fig. 7(b)), the receiving rail acts as a cantilever beam. This is because, under these circumstances, the distance between the rail end and the rail fastening system that serves as the fulcrum becomes longer, and it is thought that a large load will be placed on the rail fastening system when the load is transferring. The assumed running speed was 100 km/h. The target test load, duration of action, and impulse were 100 kN, 19 msec, and 950 N·s.
3.1.3 Test No. 3 (Curve section test to confirm rail fastening system damage)Test No. 3 was a test to confirm damage to the rail fastening system, under the condition that the rail gap begins from the edge of the sleeper on the leaving side in a curved section. This is to check the effects of the lateral force as well as the wheel load when running in curves. The assumed curve parameters were a radius of 620 m, a cant of 50 mm, and a running speed of 90 km/h. The target test load, duration of action, and impulse were 98 kN, 23 msec, and 1,127 N·s. Due to the limitations of the test device, loads cannot be applied simultaneously from two directions, so the test load in the curved condition was the combined force of the wheel load and the lateral force, and the specimen was tilted by the angle of the combined force. It should be noted that the test to confirm rail damage in curves was not included in this study because it is difficult to simulate the contact condition between the wheel flange and the rail that occurs when running in a curve. However, based on the results of Test No. 1 described below in Section 3.2, it is thought that the state of rail damage in a curve can also be roughly estimated from the experimental estimation formula.
3.2 Straight lines test to confirm rail damage (Test No. 1) [8] 3.2.1 Test methodThe test conditions and a test load example are shown in Fig. 9. Test loads were converted per single rail by dividing the value measured by the load cell by 2. To ensure that the load was applied evenly to both rails, the acceleration of each rail was measured in advance and the loading position was adjusted so that the accelerations were equal. Looking at the test load shown in Fig. 9(b), compared to the target impact load of 170 kN, duration of 8 msec and an impulse of 680 N·s, the load of the first wave was 170.5 kN, duration of 10 msec and an impulse of 852.5 N·s, which exceeded the impulse of the target load conditions. It should be noted that the second and third waves also occurred after the first wave. This is the result of the weight repeatedly bouncing up and down several times after the fall. Although it is difficult to control the rebound of the weight, the impact load of the second wave was 65.1 kN, with a duration of 11 msec and an impulse of 358.1 N·s, which was significantly lower than the target load condition’s impulse. Therefore, in this study, it was not counted as a load cycle. Additionally, to confirm the amount of plastic deformation on the top surface of the rail being evaluated, the cross-sectional shape of the rail was measured. The measurement point was set at a position 5mm from the rail end in the longitudinal direction, and the vertical deflection at the center of the rail top surface was confirmed.
Figure 10 shows the amount of plastic deformation on the rail head surface and the condition near the rail head surface after 100,000 load cycles. The results of the cross-sectional profile measurement of the rail showed that the amount of depression of the rail head due to plastic deformation was 1.2 mm after 100,000 cycles. Considering the fact that there are reports [9] that rail steps of up to 3 mm have occurred at rail joints, the amount of deformation tested is small and is not considered to immediately cause a running safety problem. No other defects or cracks in the rail were observed. Additionally, the vertical depression of the rail head surface due to plastic deformation tended to converge to a constant value after approximately 50,000 cycles. These results confirmed that although the rail undergoes some plastic deformation due to the impact load when passing over the rail gap, the amount of deformation asymptotically converges to a constant value and does not lead to failure even after approximately 100,000 load cycles. Additionally, Urakawa et al have proposed an estimation formula based on experiments [5], as shown in Equation (1), for the maximum plastic deformation of the rail head surface due to the impact load at the rail gap.
| (1) |
Z pmax : Maximum plastic deformation of rail (mm), Pa: Impact load (kN), σy: Rail yield stress (N/mm2) (reference value 450 N/mm2), α, β: Geometric shape coefficient (reference value α = 121.14, β = 231.39)
By applying the impact load of 170 kN, which is the test condition, to the above experimental formula, the estimated maximum plastic deformation was 1.05 mm. This estimation agreed with the test result of 1.2 mm with an accuracy of 87.5%. This suggests that the amount of plastic deformation of the rail head due to impact loading can be estimated.
3.3 Straight line test to confirm rail fastening system damage (Test No. 2) [10] 3.3.1 Test methodThe test conditions and examples of test loads are shown in Fig. 11. Checking the test loads shown in Fig. 11(b), it can be seen that there are more repeated impacts compared to test No. 1. This is because the rail is in a condition similar to a cantilever beam, causing elastic bending deformation of the rail when the weight collides with it. This results in a spring-like response of the weight, leading to an increase in the number of impacts. Additionally, the value of the impact load for the first wave was 117.7 kN, 21 msec, and 1,235.9 N·s, which exceeded the target load conditions of 100 kN, 19 msec, and 950 N·s. Furthermore, the impact load, duration, and impulse of the second wave were 80.2 kN, 28 msec, and 1,122.8 N·s, respectively. Since the impulse of the second wave also exceeded the target load conditions, it was counted as a load cycle. However, the impact load, duration, and impulse of the third wave were 57.9 kN, 21 msec, and 608.0 N·s, respectively, which were lower than the target load conditions. Therefore, the third wave and subsequent waves were not counted as load cycles. In this test, to verify the functionality of the rail fastening system after repeated impact loads, the fastening force of the rail fastening system was measured after the test was completed.
The appearance of the rail fastening system after 100,000 load cycles is shown in Fig. 12. As a result of the test, no visible damage to the rail fastening system was observed. However, it was confirmed that the clip constituting the rail fastening system had plastically deformed upward by approximately 1.6 mm compared to an unused clip. This is because the receiving rail was in a condition similar to a cantilever beam, with impact loads applied to the rail end. As a result, the clips of the rail fastening system, which act as fulcrums, were repeatedly subjected to an upward vertical force exceeding the rail fastening force. After the test, the rail fastening force was 21.2 kN for the two clips, with an average fastening force of 10.6 kN per clip. The result represents a reduction of 11.7% compared to the nominal rail fastening force of 12.0 kN for the spring-type rail fastening system used in this test. Regarding the performance of the rail fastening system with reduced rail fastening force, a performance verification test was previously carried out on a similar rail fastening system. In this test, the rail fastening force per clip was reduced to 9.3 kN, and the train equivalent load was repeatedly applied 1,000,000 times. The result reported that no abnormalities were observed in the rail fastening system after the test [11]. As the rail gap was not the subject of the previous study, the results of the previous study do not directly correspond to the results of the falling weight test carried out this time. However, taking into account the results of the previous study, it is believed that the functional degradation of the rail fastening system in this test does not significantly reduce the holding function of the rail.
The test conditions and examples of test loads are shown in Fig. 13. In this test, to simulate the wheel load and lateral force during the curve passage, the specimen was inclined as shown in Fig. 13(a), and the resultant force of the wheel load and lateral force were set as the impact load. Checking the test loads shown in Fig. 13(b), repeated impacts occurred after the first wave, similar to previous tests. Additionally, the value of the impact load for the first wave was 102.9 kN, 27 msec, and 1,389.2 N·s, which exceeded the target load conditions of 98 kN, 23 msec, and 1,127 N·s. Furthermore, the impact load, duration, and impulse of the second wave were 47.7 kN, 35 msec, and 834.8 N·s, respectively. Since these values were lower than the target load conditions, the second wave was not counted as a cycle. To verify the functionality of the rail fastening system after repeated impact loads, the fastening force of the rail fastening system was measured at the end of the test, similar to Test No. 2.
The insulator, which is a resin component of the rail fastening system, was damaged after 55,000 loading cycles. The appearance of the rail fastening system after 55,000 loading cycles is shown in Fig. 14. No deformation was observed in the damage process up to approximately 30,000 cycles. However, as shown in Fig. 15(a), the insulator, which is a resin component of the rail fastening system on the outer side of the track, began to crack at around 32,000 cycles, and was finally damaged at 55,000 cycles. The reason for this is that the receiving rail was in a cantilever beam-like condition, with the impact lateral force from passing through a curve acting on the rail end. As a result, the rail fastening system, acting as a fulcrum, experienced a lateral impact load. The load is transmitted to the rail, insulator, and the shoulder part that fixes the clip. Among these, the insulator, being a resin component, has the lowest strength, which likely led to the cracks and damage. Since the behavior of the falling weight was unstable, the test was terminated after 55,000 loading cycles.
The rail fastening force of the rail fastening system was then measured after the test. It should be noted that when the insulator is damaged, the deflection of the clip decreases, resulting in a reduction in the rail fastening force. Therefore, to confirm the reduction in rail fastening force due to the plastic deformation of the clip itself, excluding the effects of damaged insulators, the rail fastening force was measured using new insulators. As a result of the measurement, the rail fastening force per clip was 11.2 kN, which is a reduction of 6.7% compared to the nominal rail fastening force of 12.0 kN. However, as mentioned in Section 3.3, this reduction is not considered to significantly decrease the rail holding function. Next, considering the damage to the insulator, we examined the rail fastening force in the condition where one side of the insulator was missing. However, in the condition where one side of the insulator is missing, it is not possible to conduct a stable rail fastening force measurement test. Therefore, using the analytical model [13] shown in Fig. 16, the rail fastening force was calculated for the condition where one side of the insulator was missing. The analytical model represents the components as three-dimensional solid elements. Additionally, the clip is modeled as an elastoplastic material, and its stress-strain characteristics are determined by performing tensile tests on material specimens cut from the straight section of the clip. The rail and insulator are modeled as linear elastic bodies, while the rail pad is modeled as a nonlinear elastic body. The elastic modulus is set to obtain the nominal spring constant. In this analysis, we removed one side of the insulator from the analytical model, which was initially fastened with insulators on both sides, and calculated the rail fastening force in this condition. The physical properties used in the analysis are shown in Table 5.
| Parts | Material model | Young’s modulus (N/mm2) | Poisson ratio |
| Rail | Elastic body | 2.06×105 | 0.3 |
| Insulator | Elastic body | 3.00×103 | 0.35 |
| Shoulder/Baseplate | Nonlinear elastic body | - | - |
| Rail pad | Elastic-plastic body | Non-linear value setting based on experiments | |
| Clip | Elastic-plastic body | ||
As a result of the analysis, the rail fastening force of the clip on the side without the insulator was 3.5 kN. On the other hand, the rail fastening force of the clip on the side with the insulator was 12.0 kN, and the combined fastening force of the two clips on both sides was 15.5 kN. This is approximately equivalent to the rail fastening force of a typical plate spring type rail fastening systems [14]. Therefore, even after the insulator is damaged, although the rail fastening condition is incomplete, the rail holding function is considered to be maintained. Based on the above results, it was confirmed that the strength of the insulator, a resin component of the rail fastening system, is crucial against the impact load assumed from the repeated passage of vehicles at the rail gap in curved sections.
3.5 DiscussionBased on the results reported in the previous sections, the damage to the track components during the repeated passage of vehicles at the rail gap was clarified through experimental investigation. Specifically, in the straight section, it was confirmed that the rail and rail fastening system experienced approximately 1.5 mm of plastic deformation at the rail head and approximately 1.6 mm of plastic deformation in the clip due to the impact load of passing over the rail gap, but no significant damage was observed. Therefore, under the conditions examined in this study, it is considered that there are no strength problems with the rail and rail fastening system during the repeated passage of vehicles in the straight section. In the curved section, it was found that the insulator, a resin component of the rail fastening system, could be damaged due to the lateral force when passing over a rail gap. However, it was confirmed that immediate damage does not occur, and no problems were observed up to approximately 30,000 cycles. Therefore, under the conditions examined in this study, it is considered that the strength problems of the rail fastening system in the curved section are minimal for repeated passages up to approximately 30,000 cycles. It should be noted that these results are based on the limited conditions examined in this study. However, for other different track structures, it is considered possible to clarify the conditions using numerical analysis based on the load conditions and test results of this study.
In this study, to obtain basic insights into the damage to track components by the repeated passage of vehicles over rail gaps on ballastless track, a falling weight test device capable of simulating the impact load of a railway vehicle passing over a rail gap was fabricated. Furthermore, repeated falling weight tests were carried out to confirm the damage to track components. The main results are as follows:
(1) By using the fabricated falling weight test device, it became possible to repeatedly apply the rail gap impact load to track components.
(2) As a result of the repeated falling weight tests simulating the conditions where the rail is subjected to a load in a straight section, the plastic deformation at the rail head was approximately 1.2 mm after 100,000 cycles of repeated loading. Considering that the rail joint has a misalignment of approximately 3 mm due to the drop of the rail, this amount of deformation was small.
(3) As a result of the repeated falling weight tests simulating the conditions under which the rail fastening system is loaded in a straight section, it was confirmed that no significant external damage to the rail fastening system was observed after 100,000 cycles of repeated loading. Additionally, as a functional check of the rail fastening system after the test, the rail fastening force was measured. The results showed that although the rail fastening force was reduced by 11.7% due to repeated impact loads, the rail fastening function was maintained.
(4) As a result of repeated falling weight tests under conditions simulating the load of a passing railway vehicle on a rail fastening system in a curved section, the insulator, which is a resin component, was damaged after approximately 55,000 repetitions. However, since no problems were observed up to approximately 30,000 repetitions, it was confirmed that the rail fastening system does not cause any immediate problems.
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Katsutoshi SHIOTA
Assistant Senior Researcher, Track Structures and Components Laboratory, Track Technology Division Research Areas: Turnout, Railway Vehicle Dynamics Simulation |
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Yuki NISHINOMIYA, Ph.D. Senior Researcher, Track Structures and Components Laboratory, Track Technology Division Research Areas: Continuous Welded Rail, Railway Vehicle Dynamics Simulation |
