In recent years, the frequent occurrence of high-rise building fires, characterized by their strong destructive power and challenging rescue conditions, has resulted in significant casualties and property losses. Utilizing ropes to descend from upper-floor platforms or windows presents a relatively simple and feasible self-rescue method. Consequently, various high-rise escape devices have emerged. Since the 1980s, extensive research on building slow-descent devices has been conducted globally, primarily categorized into damping-type and speed-reduction disc-type devices. Damping-type descenders often exhibit an increase in descent speed with the user’s weight, while speed-reduction disc-type devices tend to be structurally complex, sometimes requiring motor drive or control, leading to cumbersome operation and hindering widespread adoption. An ideal high-rise slow-descent device should be safe and reliable while also ideally possessing the following characteristics.
- The descent speed should be automatically controlled. The device should adapt to individuals of different weights without requiring manual adjustment or intervention during use, while ensuring the fastest possible escape speed under safe conditions.
- It should require no professional training, special physical strength, or particular skills from the user. Operation should be simple and intuitive, allowing for self-rescue without assistance from others.
- The structure should be simple, preferably without any power-driven components.
Addressing these requirements, I have designed a novel type of high-rise slow-descent device. Through the deceleration effect of a cycloidal drive and hydraulic damping, the device completes one cycle of descent-deceleration over a short distance of 0.8 to 1 meters. This mechanism segments the long-distance descent from a high point to the ground into several short segments, achieving a graded or intermittent slow-descent purpose. Based on the deceleration principle of the cycloidal drive, the kinematic profile of the device was designed. A numerical model of the entire machine was subsequently established, and the strength of its critical components was verified using ANSYS software. Prototype testing has validated the feasibility and safety of the device.
Structural Design and Working Principle
The schematic of the device is shown in Figure 1, consisting of four main parts: a manual rope-retrieval mechanism, a rope drum, a deceleration mechanism, and a damping mechanism.
The damping mechanism employs hydraulic orifice damping. A piston inside a hydraulic cylinder is equipped with a check valve and a damping orifice. Fluid flow from the lower chamber to the upper chamber through the small orifice is subjected to damping, while flow from the upper chamber to the lower chamber through the check valve is unrestricted. The deceleration mechanism is a harmonic drive composed of a cycloidal drive. The outer contour of the pin wheel (or a cam attached to its output) is designed as a cam profile. When a weight causes the rope drum to rotate several times, the cycloidal drive reduces this motion to one revolution of the cam. During the return stroke of the cam, the piston rod separates from the cam profile, allowing the person or weight to descend freely. During the forward (or pressing) stroke, the cam presses down on the hydraulic cylinder piston, impeding the rotation of the rope drum and thereby slowing the descent speed of the person or weight until it reduces to zero. The manual rope-retrieval mechanism consists of a pair of gears with a 3:1 transmission ratio, enabling rapid retrieval of the rope. To enhance operability, two one-way clutch (overrunning clutch) structures are incorporated. The first is installed between the manual retrieval mechanism and the rope drum, allowing the handle to drive the drum during retrieval but preventing drum rotation from driving the handle during descent, thus avoiding potential injury from a spinning handle. The second clutch is placed between the rope drum and the cycloidal drive input, allowing the drum to drive the deceleration mechanism during descent but disengaging it during retrieval to reduce resistance.
Cycloidal Drive Design and Analysis
The cycloidal drive, known for its high transmission ratio, efficiency, and compact structure, has found widespread application in recent years in fields such as robotics, CNC machine tools, and automation equipment. The designed cycloidal drive reducer can be divided into three sections: the input shaft, the reduction section, and the output section. Its kinematic diagram is shown in Figure 2. When the input shaft rotates the eccentric sleeve once, due to the unique tooth profile of the cycloidal disk and the constraint imposed by the pin teeth on the stationary pin wheel, the motion of the cycloidal disk becomes a planar movement involving both revolution and rotation. This results in the output (often connected to the cam) rotating in the opposite direction by a fraction of a revolution, achieving significant speed reduction and yielding a low output speed suitable for our intermittent motion profile.
Transmission Ratio and Tooth Profile Design
Aiming for a deceleration cycle approximately every 1 meter, the rope drum’s innermost diameter is 63 mm (circumference ~197 mm) and its outermost diameter is 102 mm (circumference ~316 mm). For transmission ratio calculation, the average circumference of 263 mm is used. Therefore, the required transmission ratio is \( i = 1000 / 263 \approx 3.93 \), which is rounded to \( i = 4 \) for design.
Three-dimensional modeling of the cycloidal drive was performed using SolidWorks. Considering the complexity of the cycloidal tooth profile, conventional modeling methods are unsuitable. I employed formula-defined curves to generate the tooth profile. Using the “Formula Curve” function in CAXA electronic drawing board and modifying the traditional mathematical equations for the tooth profile to meet specific design needs, an optimized algorithm was derived. Based on the cycloidal motion law, the tooth profile curve equations were established in a Cartesian coordinate system.
For the Pin Wheel (stationary ring):
$$ X(t) = R \sin\left(\frac{t}{2Z_a}\right) – 2r \sin\left(\frac{t}{2}\right) \cos\left(\frac{t}{2} + \frac{t}{2Z_a}\right) $$
$$ Y(t) = R \cos\left(\frac{t}{2Z_a}\right) + 2r \sin\left(\frac{t}{2}\right) \sin\left(\frac{t}{2} + \frac{t}{2Z_a}\right) $$
For the Cycloidal Disk (moving element):
$$ X(t) = R \sin\left(\frac{t}{2Z_b}\right) – 2r \sin\left(\frac{t}{2}\right) \cos\left(\frac{t}{2} – \frac{t}{2Z_b}\right) $$
$$ Y(t) = R \cos\left(\frac{t}{2Z_b}\right) – 2r \sin\left(\frac{t}{2}\right) \sin\left(\frac{t}{2} + \frac{t}{2Z_b}\right) $$
Where:
\( R \) is the base circle diameter of the cycloidal disk,
\( r \) is the pin wheel base circle radius,
\( Z_a \) and \( Z_b \) are the number of teeth on the pin wheel and cycloidal disk, respectively.
Selecting pin wheel teeth \( Z_a = 12 \), pin wheel base radius \( r = 15 \) mm, and transmission ratio \( i = 4 \), the cycloidal disk teeth \( Z_b = Z_a – i = 8 \). However, for a standard cycloidal drive, \( i = Z_a / (Z_a – Z_b) \). Setting \( i=4 \) and \( Z_a=12 \) gives \( Z_b = 9 \). The base circle diameter \( R = 2rZ_b \). Substituting these into the cycloidal equations, a single tooth profile is plotted in CAXA and then arrayed to form the complete tooth set. The output section features a cam profile. For structural compactness, the outer contour of the pin wheel housing (or a separate cam attached to the output) is designed as the cam. The generated cycloidal disk tooth profile and the integrated cam contour are illustrated in Figure 3.
| Parameter | Symbol | Value | Unit |
|---|---|---|---|
| Transmission Ratio | \( i \) | 4 | – |
| Pin Wheel Teeth | \( Z_a \) | 12 | – |
| Cycloidal Disk Teeth | \( Z_b \) | 9 | – |
| Pin Wheel Base Radius | \( r \) | 15 | mm |
| Cycloidal Disk Base Diameter | \( R \) | \( 2 \times r \times Z_b = 270 \) | mm |
| Tooth Width | \( B \) | 20 | mm |
Strength Analysis of the Cycloidal Drive
The primary failure modes for the meshing surfaces of a cycloidal drive are fatigue pitting and scuffing. The contact between a pin and the cycloidal disk tooth can be approximated as the contact of two instantaneous cylinders. According to the Hertzian contact stress formula, the contact stress \( \sigma_j \) is:
$$ \sigma_j = 0.0418 \sqrt[3]{ \frac{P_i \, E_d^2}{B^2 \, \rho_d^2} } \quad \text{(MPa)} $$
Where:
\( P_i \) is the normal force at the contact point at any instant,
\( E_d \) is the equivalent modulus of elasticity for the contacting bodies (for alloy steel, \( E_d \approx 206 \) GPa),
\( B \) is the width of the cycloidal disk,
\( \rho_d \) is the equivalent radius of curvature at the contact point.
When the tooth is loaded, the bending moment at the tooth root is maximum, making it the most vulnerable to bending fatigue failure. Referencing the bending fatigue strength check formula for standard spur gears and accounting for stress concentration at the root fillet and other influencing factors, the bending strength condition is:
$$ \sigma_f = \frac{K F_t Y_{Fa} Y_{Sa}}{B m} \leq [\sigma_f] $$
Where:
\( [\sigma_f] \) is the allowable bending fatigue stress,
\( m \) is the module (note: the equivalent module for a cycloidal drive is defined differently and is typically larger than for standard gears; a value must be chosen based on the specific design and load),
\( K \) is the load factor (typically taken as 1.8-2.0 for such applications),
\( F_t \) is the tangential force on the tooth,
\( Y_{Fa} \) is the tooth form factor (approximately 2.44 for cycloidal profiles under given conditions),
\( Y_{Sa} \) is the stress correction factor for load application at the tooth tip (approximately 1.65).
For a maximum load of 200 kg (weight force \( W = 200 \times 9.81 \approx 1962 \) N) acting on the rope drum with an average radius \( r_{drum} \approx 0.0415 \) m, the torque on the drum is \( T_{drum} = W \times r_{drum} \approx 81.4 \) Nm. With a cycloidal drive reduction ratio of \( i=4 \), the input torque to the cycloidal drive is roughly \( T_{in} = T_{drum} / i \approx 20.35 \) Nm (neglecting efficiency losses for initial estimate). The tangential force \( F_t \) on the cycloidal teeth can be estimated from this torque and the operating pitch circle radius. These values are used for preliminary hand calculation before detailed FEA.
Finite Element Model Establishment and Analysis
Numerical Model and Meshing
The geometric model of the slow-descent device was created in SolidWorks and then imported into HyperMesh for meshing to establish the finite element model. Minor features with negligible impact on structural stress, such as small fillets and edge chamfers, were simplified to improve computational efficiency while maintaining analysis reliability.
Within HyperMesh, the model was discretized using tetrahedral Solid187 elements. To accurately capture stress variations, a finer mesh was applied to the cycloidal drive components and the piston rod. The total number of elements was 863,242. The overall finite element model is shown in Figure 4. The material properties assigned were as follows:
| Component | Material | Elastic Modulus (E) | Poisson’s Ratio (μ) | Density (ρ) |
|---|---|---|---|---|
| Base Plate, Side Plates, Rope Drum, Cycloidal Drive | Alloy Steel | 206 GPa | 0.3 | 7.8×10³ kg/m³ |
| Hydraulic Cylinder | Aluminum Alloy | 69 GPa | 0.33 | 2.9×10³ kg/m³ |
Analysis Results
The finite element model from HyperMesh was imported into ANSYS Workbench for structural static analysis. A fixed support constraint was applied to the base plate. The analysis simulated the stress distribution under the maximum operational load of 200 kg.
Figures 5 and 6 display the equivalent (von Mises) stress contours on the pin wheel and the cycloidal disk, respectively. The results show that the maximum contact stress occurs at the points where the tooth profiles interact. The stress on the cycloidal disk is higher than on the pin teeth, which is expected as it is the moving, load-bearing element. Critically, the maximum stress values in both components were found to be well below the yield strength of the alloy steel (typically >500 MPa), confirming that the cycloidal drive design satisfies the strength requirement for bending and contact fatigue under the specified load.
Figure 7 shows the stress distribution on the right-side support plate. The maximum equivalent stress is concentrated at the connection point between the support plate and the base plate. This stress concentration is primarily due to the bending moment induced by the hanging load acting on the support structure. The stress value, while being a local maximum, is also within the safe allowable limit for the material, indicating the overall frame design is robust. The FEA identified this as a potential area for design improvement, such as adding a gusset or increasing the fillet radius, to further enhance durability and safety.
Prototype Experiment and Performance Validation
To validate the reliability and performance of the device, physical prototype testing was conducted. The assembled slow-descent device is shown in Figure 8. Tests were performed with different weights: a 15 kg mass, a 40 kg mass, and a 64 kg person. The device was installed at a height of 6.5 meters (simulating a third floor). The total descent time and average speed for each load were recorded and analyzed.
| Load (kg) | Total Descent Height (m) | Observed Segment Length (m) | Total Time (s) | Average Speed (m/s) |
|---|---|---|---|---|
| 15 | 6.5 | ~0.8 | 7.8 | 0.83 |
| 40 | 6.5 | ~0.8 | 6.4 | 1.02 |
| 64 | 6.5 | ~0.8 | 5.3 | 1.23 |
Analysis of Table 3 leads to several key conclusions. The device successfully facilitated the slow descent of all test loads. The fundamental intermittent descent profile—acceleration followed by deceleration to a near-stop within each segment—was maintained regardless of weight, confirming the core working principle. The segment length remained approximately constant at 0.8 meters, as designed by the cycloidal drive ratio and cam profile. As expected, the average descent speed increased with load. This is attributable to the increased force on the hydraulic piston during the cam’s press stroke for heavier loads, resulting in a higher piston velocity against the fluid damping and thus a shorter deceleration phase per cycle. However, the variation in average speed (from 0.83 to 1.23 m/s) is within a reasonable and safe range for emergency egress, demonstrating the device’s automatic adaptation to user weight without requiring manual adjustments.
Discussion and Design Considerations
The successful development of this device hinges on the synergistic integration of its core mechanisms. The cycloidal drive is the heart of the motion control system. Its high reduction ratio in a compact package is essential for translating multiple drum rotations into a single, controlled cam revolution. The dynamic load sharing among multiple teeth in a cycloidal drive contributes to its high torque capacity and smooth operation, which are critical for safety. The design of the tooth profile using modified cycloidal equations ensured proper meshing and load distribution, as validated by the FEA results which showed acceptable stress levels.
The hydraulic damping system provides the velocity-dependent resistive force necessary for gentle deceleration. The use of a simple orifice for damping and a check valve for free return stroke is a passive, reliable solution. The damping force \( F_d \) can be approximated by:
$$ F_d \propto \mu \frac{A_p^2}{C_d A_0^2} v_p $$
where \( \mu \) is fluid viscosity, \( A_p \) is piston area, \( C_d \) is discharge coefficient, \( A_0 \) is orifice area, and \( v_p \) is piston velocity. This square-law relationship with velocity helps in achieving a non-linear deceleration profile that feels natural. The cam profile was designed to convert the rotary output of the cycloidal drive into the linear stroke of the piston. The shape of this profile directly influences the deceleration curve—a polynomial or modified trapezoidal profile could be optimized to minimize jerk and provide the most comfortable stop.
Material selection played a vital role. High-strength alloy steel for the cycloidal drive and critical load-bearing parts ensures durability under repeated use and shock loads. Aluminum alloy for the hydraulic cylinder housing reduces overall weight. The FEA was instrumental not just in verification, but in the iterative design process, allowing for the identification and reinforcement of stress concentration areas like the support plate connection before physical fabrication.
From a human factors perspective, the average speeds achieved (0.8-1.2 m/s) are comparable to those of trained professionals using controlled descent techniques. The intermittent stop-and-go rhythm, while different from a continuous descent, provides psychological reassurance and allows the user to reorient themselves during each brief pause. The automatic weight adaptation eliminates a critical user error point found in many manually-adjusted descenders.
Conclusion
In summary, a novel high-rise slow-descent device has been designed, analyzed, and prototyped. The key findings and contributions are:
- Innovative Motion Principle: The device successfully implements an intermittent, graded descent profile through a purely mechanical system, mimicking professional rescue techniques. This is achieved without electrical power or complex controls.
- Central Role of Cycloidal Drive: The cycloidal drive proves to be an exceptionally suitable mechanism for this application. Its compact size, high reduction ratio, inherent load-sharing capability, and robustness provide the reliable foundation for the timing and force transmission required for the intermittent descent cycle. The design and strength analysis of the cycloidal drive components were crucial to the device’s integrity.
- Automatic Adaptation: The combination of the fixed-ratio cycloidal drive and velocity-squared hydraulic damping creates a system that automatically adjusts its descent cycle time based on the user’s weight, fulfilling a major design requirement for universal usability.
- Validated Safety and Performance: Finite Element Analysis confirmed the structural strength of critical components under maximum load. Physical prototype tests with varying weights demonstrated the device’s functional operation, safe descent speeds, and automatic weight adaptation capability.
This device, utilizing the deceleration principle of a cycloidal drive and hydraulic damping to achieve intermittent, graded descent, represents a fixed-installation escape appliance suitable for public buildings. The design is compact, safe, and reliable, offering a new conceptual approach and reference value for the future development of slow-descent and emergency escape equipment. Future work could focus on further optimization of the cam and damping profile for comfort, miniaturization of the overall package, exploration of alternative materials, and long-term durability testing under various environmental conditions.