Machine Room Passenger Elevator functions
- Mature and reliable structure: The classic design has been verified for many years and is suitable for all types of buildings. It is easy to maintain and has strong versatility of parts.
- Strong load-bearing capacity: It supports a variety of load-bearing configurations from 320kg to 2000kg to meet different passenger flow requirements.
- Efficient operation system: It adopts variable frequency and voltage (VVVF) technology, which can start and brake smoothly, save energy and reduce noise.
- Independent machine room design: It is convenient for wiring, ventilation and heat dissipation, and effectively protects electrical equipment to extend its service life.
- High safety performance: It is equipped with multiple safety devices such as overspeed protection, speed limiter, buffer, emergency leveling, fire return, etc. as standard.
- Strong customization flexibility: The car decoration, control panel, floor display, etc. can be customized according to project requirements.
Machine Room Passenger Elevator manufacturer
Full industry chain integration capability
As a leading enterprise in the entire industry chain of the elevator industry, it integrates R&D, manufacturing, sales, installation and maintenance, providing one-stop solutions to ensure efficient coordination of products and services.
Intelligent and environmentally friendly product advantages
The products cover multiple categories such as passenger elevators, sightseeing elevators, and freight elevators, focusing on core features such as intelligent control, silent operation, energy saving and environmental protection, which meet the modern market's demand for green intelligent technology.
Rich experience and technical accumulation
With more than 20 years of industry experience and 3,000+ technical solutions, it has strong technical strength and can accurately respond to diverse needs to ensure product reliability and innovation.
Customized services and perfect after-sales
Provide personalized customization services, equipped with a professional team and 24-hour after-sales support to ensure that customers obtain safe, efficient and continuously optimized vertical transportation solutions.
Machine Room Passenger Elevator Main Frame Design - A Philippine Case Study
Volkspace uses a real-world case study from the Philippines to briefly explain the considerations that should be taken into account when designing elevator machine frames.
The elevator machine frame, as a critical component in geared passenger elevators, supports the combined weight (W) of the elevator car, load, counterweight system, and the mass of the motor, traction ropes, and compensation chains. It acts as a support structure between the load-bearing beams and the traction machine, requiring guaranteed strength and rigidity.
Case Study
Elevator site parameters: Geared passenger elevator with machine room, rated speed v = 1.75 m/s, rated load Q = 1,350 kg, car weight P = 1,200 kg. The traction machine and frame structure are shown in Figure 1. The traction sheave is a cantilever structure, and the frame is divided into an upper frame and a lower frame. To prevent slight tilting and shaking deformation of the traction machine frame towards the traction sheave side when the elevator car is loaded with 1.25 times the rated load, our engineering team instructed the customer to perform on-site reinforcement welding.
Calculation and Analysis Process
Wire Rope Span (lPP Value) Standard Condition
According to general design standards, the maximum lPP value is 660mm. The forces on the machine frame are calculated based on the relevant dimensions and the maximum rated load. The maximum rated load Q = 2000kg, the weight of the elevator car P = 1800kg, the weight of the wire ropes and compensation chains, etc., Msteel + Mcompensation = 300kg, the weight of the traction machine = 800kg, and the elevator balance coefficient is calculated as 0.5. The stress on the main frame under the safety brake activation condition is analyzed (assuming the safety brake impact coefficient is 2). The three-dimensional structure of the main frame is shown in Figure 3.
Calculation Data
Total mass on the car side:
G = 2 × (P + 1.25 × Q + Wsteel + Wcounterweight) × gn = 2 × (1800 + 1.25 × 2000 + 300) × 9.81 = 90,252 (N);
Therefore, the tensile force of the steel rope on the car side is T1 = G/2 = 45,126 N.
Counterweight side total mass:
W = (P + 0.5 × Q) × gn = (1800 + 0.5 × 2000) × 9.81 = 27,468 (N);
Therefore, the tension in the steel wire rope on the counterweight side is T2 = W1 = W2 = W/2 = 13,734 (N).
The traction machine weighs 800 kg, and its weight is W_traction = M_traction × 9.81 = 7,848 N. W_traction = 8,000 N. See Figure 4 for a detailed force diagram.
As shown in Figure 4, the most unfavorable force condition occurs when the angle θ approaches 0.
In this case, T2 = W1 = W2, therefore:
T1 = 45,126 N, T2 = 13,734 N, T_total = T1 + T2 = 58,860 N.
Support reaction force RA-RB: T_total × (1 + 104.5/160) + W_towing / 2 = 101,227 N.
Support reaction force RC-RD: -T_total × (104.5/160) + W_towing / 2 = -34,537 N.
Where: RA, RB, RC, and RD are the support reaction forces at points A, B, C, and D on the frame, respectively.
Note: The positive and negative signs indicate opposite directions of the forces.
Calculation of the bending moment generated by the force on the tractor frame:
Using a value of 660 mm for lPP, we get θ ≈ 10°, and W1×sinθ = 2384 N.
As shown in Figure 5, the bending moment generated by W1×sinθ (clockwise direction) is:
MW1×sinθ = W1×sinθ×382.5 = 911,880 (N·mm).
The bending moment generated by T1 is:
MT1 = T1×(480/2) = 10,830,240 (N·mm);
The bending moment generated by T2 is:
MT2 = T2×(480/2) = 3,296,160 (N·mm).
As shown in Figure 6, the difference in bending moment between T1 and T2 (counterclockwise direction) is
MT1-T2 = MT1 - MT2 = 10,830,240 - 3,296,160 = 7,534,080 (N.mm).
The difference between MT1-T2 and the bending moment MW1 × sinθ is
M = MT1-T2 - MW1 × sinθ = 7,534,080 - 911,880 = 6,622,200 (N.mm).
The analysis shows that the bending moment produced by T1 and T2 is much larger than the bending moment produced by W1×sinθ, meaning the resulting bending moment from the difference between the two is in the counterclockwise direction.
As shown in Figure 7, the force generated by the bending moment M on the main frame is
f = 6,622,200 / 560 / 2 = 5,913 (N).
The forces at points A, B, C, and D are calculated as follows:
RA = (RA - RB) / 2 + f = 101,227 / 2 + 5,931 = 56,527 (N);
RB = (RA - RB) / 2 - f = 101,227 / 2 - 5,913 = 44,701 (N);
RC = (RC - RD) / 2 + f = -34,537 / 2 + 5,913 = -11,356 (N);
RD = (RC - RD) / 2 - f = -34,537 / 2 - 5,913 = -23,182 (N).
Based on the above analysis, it can be concluded that within the normal lPP range, the deformation and stress of the upper and lower frames of the traction machine under the safety brake activation condition meet the requirements, as shown in Figures 8 and 9.
Non-standard IPP values
According to market research statistics, in machine-room passenger elevators with a load capacity of 1,250 to 1,350 kg, when the counterweight is positioned at the rear, the following occurs: when a single sheave is configured on the car top, the lPP value increases, leading to an increase in the width of the lower frame, thus affecting the overall deflection of the entire frame.
Based on the aforementioned field case, with a load capacity of 1350 kg and a counterweight positioned at the rear, the lPP value reached 840 mm, representing a non-standard design structure. Furthermore, to meet the traction force requirements, additional counterweight blocks were added to the car bottom, resulting in increased self-weight. The forces at the corresponding points, calculated using the method described above, and the results obtained through software analysis are shown in Figures 10 and 11.
Analysis shows that increasing the lPP value results in the overall strength of the frame meeting the requirements, but the overall deformation increases. This is because increasing the lPP value increases the width of the lower frame, affecting the overall bending resistance (see Figure 12).
Response plan
Regarding the issues mentioned above, the non-standard increase in the lPP value leads to an increase in the lRL value of the lower frame. In Figure 13, the arrows indicate an increased span of the channel steel, resulting in increased deformation, which in turn increases the deformation of the upper frame. To address this structural issue, the channel steel at the position indicated by the guide lines in Figure 14 is reoriented so that the larger surface of the channel steel faces outwards. The software analysis results for this modified structure are as follows:
Applying the forces calculated in Section 2.2 of this paper to the model, the analysis results are shown in Figures 15 and 16. It can be seen that the modified structure significantly improves deformation resistance. It is recommended that non-standard structures be designed according to this approach.
Conclusion
Based on the data analysis above, to prevent slight shaking of the frame caused by the widening of the lower frame leading to deformation of the upper frame, this structure can be adopted while ensuring convenient installation of components on site. The standard product can be designed according to the structure shown in Figure 14, which not only significantly improves strength but also provides excellent resistance to deformation; furthermore, it does not increase any costs compared to the original structure. Most importantly, it maximizes the lPP value while meeting the traction force requirements.
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