Vibration control and structural optimization of steering brake assist device in new energy buses

Abstract

The aim of this study was to explore the vibration challenges caused by the crankshaft design in the air pump of the power integration device for steering and brake assistance in new energy buses. Three distinct crankshaft structures were designed and transient dynamic analysis was conducted using ANSYS software to assess the stress conditions and obtain the torque curves at the center points of these structures. Three newly fabricated crankshaft prototypes were subjected to vibration intensity tests on the integrated device. The experimental results revealed that the assistance device using the new crankshaft significantly reduced vibration intensity.

Keywords:

• New energy buses
• steering assistance
• brake assistance
• crankshaft
• vibration intensity

1. Introduction

The steering assistance system [1] and brake assistance system [2,3] are vital components in automobiles, enhancing safety and comfort by reducing the effort required for steering and braking while driving. This improves vehicle controllability and safety. Hydraulic power steering systems are commonly used in steering assistance systems, whereas vacuum servo brake systems are used in brake assistance systems. In modern vehicles, the auxiliary functions of steering and brake systems are primarily controlled electronically, leading to improved efficiency, response speed and integration with other vehicle safety systems. With the emergence of new energy vehicles and the advances in intelligent driving technologies, highly integrated systems play a significant role in enhancing vehicle performance and driving experience [4]. These systems conserve energy and effectively reduce automobile costs and weight.

A novel power integration device for steering and braking was designed for new energy buses, as illustrated in Figure 1. The primary functions of the device include providing steering assistance and brake assistance for new energy buses. From left to right, the components of the device comprises a vane-type steering assist pump [5], a high-low pressure dual-source electric motor, a sensor and control interface unit, an electromagnetic clutch, a piston-type brake assist pump, an air-cooled radiator fan, and a heat dissipation guide cover. The steering assist pump is functions through hydraulic system, whereas the brake assist pump is powered by a pneumatic system. The high-low pressure dual-source electric motor can simultaneously power the steering assist oil pump and the brake assist air pump. The steering and braking assist devices use advanced technologies [1, 6–8]. Seamless integration of these two systems reduces component costs and reduces the weight of the vehicle.

Figure 1. Integrated device for steering and braking power of new energy buses.

Preliminary testing showed that the integrated assist device exhibited significant vibration [9], primarily attributed to imbalance in the crankshaft of the piston-type air pump during rotation. The uneven mass distribution of the crankshaft and its associated components such as pistons and connecting rods generates a centrifugal force during rotation, leading to vibration. This vibration is subsequently transmitted to the motor and oil pump through the electromagnetic clutch. Therefore, optimizing the design of the crankshaft is essential for reducing the vibration [10,11] in the assist device.

During operation, the crankshaft is subjected to intricate torques and bending stresses from gravitational forces and the imbalance of inertial forces caused by its rotation [12]. Furthermore, these stresses originate from inertial forces from connecting rods, pistons, and support reactions, which vary with the crankshaft’s rotation. Additionally, bearing deformation, shaft system deformation, and wear on the crankshaft journals introduce additional loads on the crankshaft’s operation. Traditional calculation methods have several limitations due to the complex shape of the crankshaft and varying loading conditions. Scholars have extensively explored the crankshaft design [13–17], mainly focusing on analyses of crankshaft dynamic characteristics, vibration mode identification, developing vibration control technologies, and optimizing materials.Liu et al. [18] studied the torsional vibration characteristics of the shaft system of high-speed and high-power reciprocating compressor. He pointed out that the first three modes affected the torsional vibration response, and proposed a method to suppress the torsional vibration amplitude. Yingkui et al. [19] conducted finite element analysis to establish stress variation models for diesel engine crankshaft and identify the most critical stress points, leading to optimization of the crankshaft design. Zhou [20] integrated the static and dynamic characteristics of the crankshaft and used optimization methods such as BP neural networks and genetic algorithms to optimize the parameters of the crankshaft, reducing angular vibration, maximum stress values, and the weight. Thirunavukkarasu [16] achieved a 22% reduction in crankshaft weight by implementing modifications such as hollowing the crankpin and shaft diameters, while simultaneously doubling the safety factor. This approach was effectively implemented to light commercial vehicles requiring increased power in addition to the torque requirements.

Three distinct crankshaft structures were designed (referred to as Structure 1, Structure 2, and Structure 3) to optimize the vibration performance of the steering and braking device (Figure 2). Transient dynamic analyses of the three novel crankshaft structures and the original structure were conducted using ANSYS software. In addition, the contributions of the structures to vibration control in actual applications were evaluated.

Figure 2. Four different crankshaft structural models.

2. Model construction

2.1. Modification of the finite element model

A meticulous balance was maintained between the computational load and accuracy during the computational analysis of the crankshaft structural models. Some non-critical structures or components within the model, including connecting rod bolts, washers, snap rings on pistons, and piston pins, were simplified[21] to reduce the computational load without reducing the accuracy.

2.2. Mesh generation

In finite element analysis, mesh generation plays a crucial role in ensuring computational accuracy and efficiency, directly influencing the results of the analysis of the shaft system model. In this study, a free meshing method was adopted to accommodate the intricate geometric shapes of the shaft system structures. This technique enables application of different mesh types and densities in various parts of complex geometries to effectively conform to the specific shapes of each component. In critical regions of the model, such as connecting rods, crankpins, and transitions to the main bearing journals, the mesh was locally refined to ensure the required analysis accuracy without significantly increasing the computational burden. Figure 3 illustrates the finite element mesh model of crankshaft system structure 1, demonstrating the meshing strategy adopted to balance the requirements for analytical precision and computational efficiency.

Figure 3. Finite element mesh model of crankshaft structure 1.

2.3. Determination of boundary conditions

The four different crankshaft systems were analyzed through simulations. It was imperative to standardize the configuration of boundary conditions to ensure consistency between the simulation models. Additionally, some equivalent treatments were implemented to harmonize the linkage relationships and boundary conditions, minimizing the discrepancy between the simulation results and experimental outcomes. Based on the actual motion characteristics of the crankshaft system, the application of rotational constraints primarily involved two aspects: firstly, fixing both ends of the crankshaft to simulate the support conditions experienced during actual operation; secondly, applying a rotational angular velocity of 8 rad/s to the crankshaft to replicate its rotational movement. High-pressure pistons and low-pressure pistons were established with vertical moving pairs, and a vertical downward force of 1000 N was applied to the upper surface of the pistons to simulate the pressure conditions experienced during actual operation. Rotational pair constraints were applied to the connecting rods and piston pins to ensure their rotational motion matched the actual operation scenario. The specific constraints are presented in Figure 4.

Figure 4. Constraint model of crankshaft system.

3. Analysis of the computational results

Transient dynamic analysis of the crankshaft structure was performed using Ansys Workbench software. The total calculation time was set to 1 s and a time step set at 0.004 s to ensure adequate resolution accuracy. The crankshaft rotated approximately 458 degrees, exceeding a complete rotation cycle. A contour map of equivalent stress distribution was recorded at intervals of every 90° rotation of the crankshaft to comprehensively evaluate the stress distribution of the crankshaft. Figure 5 illustrates the contour map of equivalent stress distribution for the conventional crankshaft structure 0. The maximum equivalent stress value in the crankshaft structure was 21.996 MPa, with relatively uniform forces on the two connecting rods (Figure 5). However, a stress concentration phenomenon was observed at the shaft shoulder positions, particularly evident at the left end face shaft shoulder, indicating potential vulnerabilities in the structure.

Figure 5. Cloud chart of equivalent stress distribution of existing crankshaft structure 0.

The stress of the conventional crankshaft structure 0 was evaluated through time-domain signal extraction [22] (Figure 6). The analysis revealed that throughout the entire motion cycle, the minimum stress values of the crankshaft system fluctuated within the range of 1 × 10^-5 MPa to 1.5 × 10^-5 MPa, whereas the maximum stress values ranged between 13.062 and 21.996 MPa. This indicates the dynamic stress characteristics experienced by the crankshaft during operation, providing essential data for evaluating its structural performance and durability.

Figure 6. Time-domain distribution of equivalent stress for the conventional crankshaft structure.

During the rotation of the crankshaft, the force exerted by the piston on the crankshaft varies at different stages of its cycle, leading to torque fluctuations that cause vibration [23,24]. Figure 7 illustrates the torque variations at the center point of the crankshaft in the X, Y, and Z directions, revealing a direct correlation between the torque magnitude at the crankshaft’s center point and vibration levels. The results showed that the maximum torque on the conventional crankshaft reached 5.79 × 10^5 N·mm and exhibited significant fluctuations. This torque variability directly influence crankshaft vibration. These findings are consistent with current experimental observations, validating the theoretical analysis of the impact of torque fluctuations on crankshaft vibration.

Figure 7. Torque curves in the X, Y, and Z directions at the center point of the existing crankshaft.

Figure 8 shows a contour map for the equivalent stress distribution for crankshaft structure 1, which is an improved design of the crank arm built based on the conventional foundation to enhance stress conditions. The results indicated that the optimized crankshaft structure exhibited a maximum equivalent stress of 22.74 MPa, located at the main bearing shoulder position.

Figure 8. Cloud chart of equivalent stress distribution of existing crankshaft structure 1.

The torque variations at the center point of crankshaft structure 1 in the X, Y, and Z directions are presented in Figure 9. The maximum torque in model 1 was 2.37 × 10^4 N·mm, which was significantly lower compared to the conventional model. The design modifications effectively reduced the torque fluctuation range, ultimately decreasing the vibration caused by torque variations. This provides valuable insights for the dynamic optimization of the crankshaft system.

Figure 9. Torque curves in the X, Y, and Z directions at the center point of crankshaft model 1.

Figure 10 illustrates the contour map of equivalent stress distribution for crankshaft structure 2, with balance weights integrated into the crankshaft to counterbalance the centrifugal forces generated during rotation. The maximum equivalent stress on the crankshaft was 25.736 MPa (Figure 10). However, analysis of the torque variation curves at the crankshaft’s center point in the X, Y, and Z directions (Figure 11) showed that the maximum resultant torque was 4.06 × 10^4 N·mm, with relatively stable torque variations. These results indicate that integrating balance weights into the crankshaft effectively mitigates the centrifugal forces during rotation and partially stabilizes torque fluctuations, thereby reducing vibration and enhancing the overall performance of the crankshaft system.

Figure 10. Cloud chart of equivalent stress distribution of existing crankshaft structure 2

Figure 11. Torque variations in the X, Y, and Z directions at the center point of crankshaft model 2.

Figure 12 shows the contour map of equivalent stress distribution for crankshaft structure 3, with high-pressure pistons hollowed out to improve the mass distribution of the pistons, enhancing the dynamic performance of entire crankshaft system. Stress analysis indicated that the maximum equivalent stress in the shaft system was 22.618 MPa, exhibiting force characteristics similar to the conventional model. Further evaluation of the torque variation curves at the crankshaft’s center point in the X, Y, and Z directions (Figure 13) revealed that the torque fluctuations were similar to the fluctuations of the conventional model, with a slight reduction in maximum torque to approximately 4.88 × 10^5 N·mm. These results indicate that optimizing the mass distribution of the pistons affected the dynamic response of the crankshaft system, mainly reflected by a slight decrease in maximum torque. However, the overall force conditions and torque fluctuations were relatively similar to those of the existing model, indicating a limited impact on the overall force and torque fluctuations.

Figure 12. cloud chart of equivalent stress distribution of existing crankshaft structure 3

Figure 13. Torque curves in the X, Y, and Z directions at the center point of crankshaft model 3.

The analysis revealed that the location and magnitude of the maximum stress points on the crankshaft varied over time during dynamic operation (Figure 14). Stress concentration commonly occurs at the crankshaft main journal shoulder positions. Therefore, an extensive evaluation was conducted on the time-domain distribution of the maximum equivalent stress at the left-end face shoulder fillet positions of the four crankshaft system structures (Figure 15). The results showed that maximum stress values exhibited minimum variations among the four models, indicating relative consistency in stress distribution across the models (Figure 16). However, evaluation of the frequency of alternating stress revealed that crankshaft system structure 1 exhibited more rational design features. This implies that this model can effectively handle stress variations during motion, thereby enhancing the overall performance of the system.

Figure 14. Left-end Face shoulder fillet.

Figure 15. Time-domain distribution of the maximum equivalent stress at the fillet of the left-end face shoulder for the four crankshaft structures.

Figure 16. Experimental site.

4. Experimental validation of the model

Three new crankshaft structures were fabricated and compared with the conventional model using the same steering and brake assist device by conducting experimental tests to validate the accuracy of the simulation design. The tests were conducted on a specialized automotive power steering oil pump test bench (model S0117-2347). The intensity of composite vibration was measured using the effective value of vibration velocity [25] to evaluate the effectiveness of the experiment. The experiment was carried out on an assist device equipped with shock-absorbing pads, with the motor operating at 1500 revolutions per minute and the air compressor pressure set at 1 MPa. The vibration intensity of the steering assist device was tested under no-load, half-load, and full-load conditions. The experimental setup and the layout of measurement points are shown in Figures 16 and 17, respectively. The experimental setup comprised 13 test points for comprehensive analysis of vibration intensity [26].

Figure 17. Layout diagram of vibration intensity measurement points.

Table 1 shows the vibration intensity recorded at different measurement points under no-load, half-load, and full-load conditions for the assist devices equipped with the four different crankshaft structures. The device equipped with crankshaft model 1 exhibited a significant reduction in vibration compared with the one with the conventional mode, whereas the devices with crankshaft models 2 and 3 showed a slight decrease in vibration, consistent with the simulation results.

Table 1. Vibration intensity at different measurement points under no-load, Half-load, and Full-load conditions

 unit:mm/s.

Test points Model/working condition		1	2	3	4	5	6	7	8	9	10	11	12	13
Existing models 0	No load	16.9	35.9	40.6	25.7	28.3	25	9.7	3.7	18.4	15.8	25.3	93.3	22.3
	half load	17.6	38.2	47.7	28.1	30.5	26.4	9.8	5.2	16.8	18.1	35.3	95.4	22.4
	Full load	18.4	53.4	59.2	30.8	30.7	30.6	10.2	7.5	19.3	20.9	43.7	96.6	24.8
New model 1	No load	13.5	31.2	36.2	23.2	26.1	22.4	7.9	3.2	15.2	13.7	23.2	89.9	19.2
	half load	14.2	35.2	44.4	25.9	26.2	20.2	8.2	5.1	15.1	14.5	31.4	90.2	18.8
	Full load	15.8	47.4	56.2	28.2	28.1	29.3	9.2	6.9	17.2	19.3	39.2	92.1	22.9
New model 2	No load	15.8	34.5	39.9	25.2	27.2	23.6	8.8	3.7	18.2	14.6	24.8	92.2	21.2
	half load	15.1	37.2	46.3	27.4	29.2	25.1	9.1	4.9	17.9	17.2	33.2	93.4	21.9
	Full load	17.9	49.9	57.2	30.7	30.4	29.7	9.8	7.4	18.4	20.2	40.2	93.8	23.6
New model 3	No load	16.2	35.6	39.5	25.7	28.2	25.1	10	3.8	17.4	14.5	24.8	90.6	21.2
	half load	15.3	38.4	46.7	28.3	30.4	24.5	9.7	5.2	16.7	18.1	35.3	92.5	22.5
	Full load	18.2	53.4	58.8	30.4	30.4	30.2	10.1	7.3	20.3	20.1	41.1	94.3	23.4

Multiple factors were comprehensively explored during the analysis of device vibration, leading to identification of measurement points 9, 10, and 13 as the three most sensitive positions to vibration response. Therefore, these three measurement points were essential in monitoring and analyzing the vibration performance of the device, providing key information for further vibration control and optimization. For instance, under the full-load condition, the average vibration intensity at measurement points 9/10/13 of the conventional crankshaft model was 21.7. Conversely, the average vibration intensity for crankshaft model 1 at the same three points was 19.8, indicating a reduction in vibration intensity of 8.8% compared to the original model. Similarly, models 2 and 3 exhibited 4.6% and 1.8% decrease in vibration intensity compared with the original crankshaft model (Figure 18).

Figure 18. Scatter diagram of vibration intensity of four structures under full load condition.

5. Conclusion

The crankshaft was subjected to alternating stress loads during its motion. The maximum and minimum stress values and their locations across the entire shaft system consistently changed under the same boundary conditions and external loads, causing challenges in implementing the traditional design methods and extending the design cycles. The analysis of the transient dynamics for the four distinct shaft system structures, through software simulation and experimental tests, provides a time-domain analysis of the stress values across the entire shaft system and critical points on the shaft shoulders. In addition, it provides details on the torque at the crankshaft center point. Comprehensive analysis showed that the new crankshaft structure 1 experienced smaller maximum stress under the same external load conditions, resulting in reduction in vibration by 8.8% compared to the original model. These research results reveal the stress response characteristics of each crankshaft structure under dynamic conditions, highlighting the importance of structural design in enhancing stress distribution and enhancing system stability. These insights provide a valuable reference for the design and optimization of crankshaft systems.

Disclosure statement

No potential conflict of interest was reported by the author(s).

Additional information

Funding

The authors wish to thank the experimental inputs provided by Major scientific and technological projects in Wenzhou (ZG2022025) and 2023 Annual University Domestic Visiting Engineer "School-Enterprise Cooperation Project"(FG2023035).