Gas stove process valve control system

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The control system for the process valves of gas stoves was changed from hydraulic automatic control to hydraulic microcomputer automatic control in 1993, and the transformation was completed entirely by 1998. The overall operation has been satisfactory, bringing real benefits to the enterprise. The hydraulic system consists of components such as a programmable logic controller (abbreviated as PLC), solenoid valve stations, oil pump stations, accumulators, hydraulic cylinders, valve position detection feedback devices, and process valves. The operating sequence of the hydraulic system is as follows: The CB-50 gear pump supplies the hydraulic fluid for the system. After the pressure of this fluid is adjusted using a relief valve, check valve, and throttle valve, it enters the high-pressure oil main pipe. Accumulators are installed on this high-pressure pipe to compensate for any sudden changes in flow rate within the system. The hydraulic fluid then flows to the solenoid valve stations associated with each furnace, where the switching of the process valves is achieved through the rotation of the solenoid directional control valves. The entire hydraulic system utilizes a differential circuit; signals from the PLC cause the electromagnetic directional control valve to be energized or de-energized, thereby enabling it to switch directions and driving the hydraulic cylinder to move, thus achieving the direction change of the process valves. The microcomputer automatic control device for gas generation oil pressure has been in use since its introduction, and its overall operation has remained basically stable. However, as the inner diameter of the gas furnace chamber increased from Ф2745mm to Ф3000mm, and then further to Ф3300mm, and especially following the modification of the gas generator chassis in March 2002, the process valves were adjusted accordingly, while the hydraulic system remained unchanged. As a result, the opening and closing speed of process valves with a diameter of DN600 mm or larger became significantly slower, which not only posed safety risks to production but also reduced the effective time for gas production. Since the successful deployment of the furnace condition optimization system in August 2004, the production capacity of the gas furnace has been significantly increased. In particular, the slow speed of direction change for process valves has restricted the full utilization of the production capacity of gas generators. Optimization and renovation of the hydraulic system are imperative. The main problems existing in the gas generation oil pressure system can be summarized as follows: (1) The improper arrangement of system components results in a slow speed of valve direction changes, leading to inaccurate cycle times and significantly affecting the gas production volume ； (2) The cooling effect of the oil pressure system’s cooling device is poor, resulting in unstable system pressure. 1 Analysis of the reasons for the optimization and modification of the hydraulic system 1.1 Analysis of the reasons for the slow direction change of process valves (1) It is due to unreasonable equipment configuration: The hydraulic system used in our company is designed for gas generators with a diameter of Ф2745mm; the effective diameter of the solenoid directional control valve in this hydraulic system is Ф9mm, and the effective diameter of all hydraulic pipelines is also Ф9mm. The blowdown valve of the F2745mm gas generator was originally DN600mm, the tee valve was originally DN750mm, and the chimney valve is DN900mm. As the inner diameter of the gas stove furnace was increased from Ф2745mm to Ф3000mm, and then from Ф3000mm to Ф3300mm, the process valves were modified accordingly: the blowing valve was changed from DN600mm to DN750mm, with its stroke being adjusted from 450mm to 560mm; the tee valve was changed from DN750mm to DN900mm, with its stroke changing from 630mm to 880mm; the DN900mm chimney valve was equipped with a lifting-type seal, and its stroke was changed from 530mm to 900mm. The hydraulic cylinders for these three process valves were also changed from D80 cylinders (i.e., cylinders with a diameter of 80mm) to D90 cylinders. The increased travel distance of the hydraulic cylinder results in a longer movement time, which causes the process valve to reach its position more slowly ； Due to the increased cylinder diameter, there is a greater loss in hydraulic oil flow, which reduces the operating speed of the cylinder and results in slower positioning of the process valves. As the weight and outer diameter of the process valve spool increase, the area exposed to the action of the fluid also increases, leading to higher resistance. (2) Improper arrangement of hydraulic components: The lack of a suitable selection of hydraulic components results in a flow-limiting effect. It is mainly reflected in the following aspects: ① The hydraulic cylinder of the process valve has been changed from a DN0mm cylinder to a DN90mm cylinder. Due to the increased cylinder diameter, the flow rate of hydraulic oil decreases, whereas the solenoid directional valve in the hydraulic pressure system and the effective diameter of all hydraulic pipelines remain at Ф9 mm. The effective flow area does not meet the requirements, resulting in flow restriction. ②Irregular manufacturing of hydraulic components: According to GB786-76, the diameter of the inlet and outlet ports of D90 hydraulic cylinders should be Ф15mm. However, in many cases the actual effective diameter of such cylinders is only Ф10mm. As a result, the flow loss that occurs when the cylinder changes direction is not compensated for in a timely manner, which limits the speed at which the cylinder can move. ③The processing errors of the solenoid valve station integration blocks are large; the effective diameter of some A and B ports is around Ф6 mm, **which reduces the effective flow area and leads to flow restriction. (3) During the direction change of the hydraulic cylinder, both the rod side chamber and the rodless side chamber are filled with high-pressure oil; due to the small pressure difference, a certain amount of resistance is generated, which limits the speed of direction change. 1.2 Analysis of the reasons for poor cooling performance of the hydraulic system cooling device: The cooling device in our workshop’s hydraulic system is an internal coiled coil type, with cooling water pipes having a diameter of DN20 mm. Calculations show that the heat exchange area of the serpentine coil is only 1.6 m2. The heat exchange area of the cooling coiled coil in the fuel tank is too small. In summer, the oil temperature in the hydraulic system rises to 52–55°C, whereas the acceptable operating range for this temperature is 35–40°C. Excessively high oil temperatures not only accelerate the deterioration of the oil but also reduce its viscosity ； Increasing the leakage of a gear pump has a direct impact on its proper operation, making it difficult to maintain stable system pressure. 2 Optimization and modification of the hydraulic system 2.1 Flow rate design The main factor affecting the slow direction change of process valves is the system’s flow rate. According to the equation of fluid continuity: q = VA, it can be seen that when a hydraulic cylinder is in operation, the speed of the piston’s movement is equal to the average flow velocity of the liquid inside the cylinder. When the effective area of the hydraulic cylinder remains constant, the speed of the piston’s movement is determined by the flow rate of fluid supplied to the cylinder. That is, the greater the flow rate of the input or output hydraulic cylinder ； The faster the piston moves, the shorter the opening and closing time of the process valve. The DN750mm blow-off valve has a stroke of 560mm; the DN900mm tee valve has a stroke of 880mm; the DN900mm chimney valve has a stroke of 900mm. The hydraulic cylinders for these three process valves are of the D90 type. The faster the operating speed of process valves, the more beneficial it is for production. In actual manufacturing, the speed at which the cylinder of a DN900 mm chimney valve can change direction is particularly crucial. Therefore, the flow rate of the hydraulic system will now be calculated using a chimney valve with a diameter of DN900 mm and a stroke length of 900 mm. It is known that the rated flow rate of the electromagnetic directional control valve currently used in our workshop is Q = 35 L/min (which is equivalent to 0.000583 m3/s). The piston rod diameter of the D90 hydraulic cylinder is 45 mm; thus, the operating speed of this cylinder when lifting something is: V = Q/A = 0.000583 ÷ π/4 × (0.092 – 0.0452) = 0.122 m/s. The time required for the cylinder to complete a stroke of 900 mm is therefore: t = s/V = 0.9 ÷ 0.122 = 7.37 seconds. Due to factors such as the upward thrust exerted by the gas in the system and the frictional forces associated with the valve seals, which cannot be precisely calculated, the actual operating time for the 900 mm chimney valve is on average 5.6 seconds. That is, there is a deviation of about 1.8 seconds between the theoretical running time and the actual running time. To advance the actual average operating time of the cylinder from 5.6 s by 1.5 s, i.e., to 4.1 s. This is achieved by increasing the flow rate of the electromagnetic directional control valve, while keeping the system operating pressure constant. The rated flow rate required for the electromagnetic directional control valve is: Q = VA = s/t × A = 0.9/4.1 × π/4 × (0.092 – 0.0452); thus, Q = 0.001047 m3/s, which is equivalent to 62.8 L/min. By referring to the \"Mechanical Design Handbook,\" an electromagnetic directional control valve with a nominal flow rate of 63 L/min was selected as the design flow rate for the modification of the hydraulic system, and its nominal diameter is DN16 mm. Since V = Q/A, the flow rate Q is inversely proportional to the cross-sectional area A; that is, with the system oil pressure remaining constant, a smaller cross-sectional area A helps to increase the operating speed. Since the DN900 mm chimney valve is the heaviest among the process valves and presents the greatest system resistance, the minimum diameter D of the required cylinder is determined through calculations: the schematic diagram of the cylinder operation is shown in Figure 1. file:///E:/DOCUME~1/zzjjaa/LOCALS~1/Temp/ksohtml/wps_clip_image1.jpg Weight of the chimney valve stem: G1 = 250 kg; Weight of the chimney valve shaft: G2 = 80 kg ； Oil pressure system pressure: P = 4 MPa; Gas system pressure: P1 = 0.015 MPa (maximum value measured). To enable the valve to descend, the minimum value of D1 for the cylinder required is as follows: The minimum returning force of the cylinder is given by F = P1×π/4×0.92 – (G1 + G2)×9.8. Calculating this gives F = 15000×π/4×0.92 – (250 + 80)×9.8 = 9537.75 – 3234 = 6303.75 N. If the diameter of the piston rod of the hydraulic cylinder is set at d = 45 mm, then the minimum value of D1 required is: P = F/A → A = F/P. Therefore, πD1²/4 = F/P, which implies that D1² = 4F/πP. Hence, D1 = 44 mm. To lift the valve, the minimum return force required of the hydraulic cylinder is:
F = (G1 + G2) × 9.8 + P2 × π/4 × 0.92
F = (250 + 80) × 9.8 + 15000 × π/4 × 0.92
F = 3234 + 9537.75 = 12771.75 (N)

Given that the diameter of the piston rod of the hydraulic cylinder is d = 45 mm, the minimum value of D2 required is:
P = F/A → A = F/P
Thus, π/4(D2² – d²) = F/P
D2² = 4F/πP + d²
D2 = √(4F/πP + d²) = 0.078 m, or 78 mm. Refer to the Mechanical Design Handbook and select the required cylinder as the D80 cylinder. Obviously, D1=44mm < D2=78mm, which does not meet the requirement for the minimum thrust necessary for the cylinder to function; therefore, it was decided that the cylinder required for the optimization of the hydraulic system should be the D80 cylinder. 2.2 Component Matching Design (1) Design of the solenoid valve station integration block: The electromagnetic directional control valve, for which Q = 63 L/min is determined based on the aforementioned flow rate calculations, serves as the designed flow rate for the modification of the hydraulic system; its nominal diameter is DN16 mm. To meet the requirement of a nominal diameter of DN16 mm for the hydraulic system, the “P” and “T” ports of the integrated block’s oil circuit have been changed from Ф20 mm to Ф30 mm, while the “A”, “B” and “O” ports have been changed from Ф9 mm to Ф16 mm. It also improves the machining precision level, ensuring that the effective diameter of ports “A” and “B” is ≥ Ф16mm. (2) Pipeline joint design and selection: To meet the nominal diameter of DN16 mm for the hydraulic system, all original joints with a diameter of Ф10 mm were replaced with joints having a diameter of Ф16 mm, and the high-pressure hoses were changed from type Ф10 mm to type Ф16 mm. (3) Selection of hydraulic cylinders: The hydraulic cylinders for the DN750 mm air blowing valve, DN900 mm tee valve, and DN900 mm chimney valve will be changed from D90 cylinders to D80 cylinders. An increase in the operating speed of the cylinder leads to greater impact on the valves. To reduce the impact on the valve spools and seats, D80 cylinders equipped with buffers are used, and it is ensured that the effective diameter of the inlet and outlet connections of the cylinder is ≥ Ф16mm. To prevent the cylinder from coming loose frequently, the connection between the cylinder body and the flange has been changed from a threaded connection to a welded flange connection. (4) Oil cooler design: The cooling device for the oil pressure system in my workshop is an internal coiled coil, with the cooling water pipe having a diameter of DN20 mm. Calculations show that the heat exchange area of the serpentine coil is only 1.6 m2. Actual measurements indicate that the temperature of the oil after cooling is around 4°C; in summer, the oil temperature in the hydraulic system rises to 52–55°C, resulting in significant fluctuations in the hydraulic system. The normal oil temperature range is 30–40°C. To ensure a cooling temperature difference ΔT of 12–15°C for the oil, I designed and built a vertical multi-tube heat exchanger with a heat exchange area of 10 m². And the inlet and outlet pipes for the pump station cooling water were changed from DN20mm to DN32mm. 3 Optimization and modification measures for the hydraulic system: In response to the main problems existing in the system, the following modifications are carried out: (1) Replace the solenoid valve station’s integrated block with an effective diameter of Ф16mm ； (2) Replace the joint with an original diameter of Ф10mm with one having a diameter of Ф16mm ； (3) Change the high-pressure hose from Ф10mm to Ф16mm; use straight pipes for the connections at the integrated block terminals in order to minimize resistance caused by bends in the oil circuit ； (4) The hydraulic cylinders shall be D80 cylinders with buffers; the cylinders for the DN750 mm blowing valve, DN900 mm tee valve, and DN900 mm chimney valve shall be changed from D90 cylinders to D80 cylinders. Ensure that the inlet and outlet connections of the cylinder have an effective diameter of Ф16 mm, and change the connection method between the cylinder body and the flange from threaded connection to welded flange connection ； (5) The HB20-B15G24-L solenoid valve is selected, which is a new type of high-flow solenoid valve with an effective diameter of Ф16mm. Replace all the inlet and return needle valves of the hydraulic cylinders with ball valves (model: Q21H-100DN20) ； (6) An external vertical shell-and-tube heat exchanger with a heat exchange area of 10 m2 is added to the return pipeline of each oil pump station, and the inlet and outlet pipes for the cooling water in the pump station are changed from DN20mm to DN32mm. 4 Investment Costs and Benefit Analysis: In January 2005, taking advantage of the maintenance of Gas Furnace No. 14, a comprehensive renovation and installation were carried out on this furnace. The investment costs are as follows: (1) Replacing the solenoid valve station integrated module costed 0.5 million yuan ； (2) Replacing all joints with a diameter of Ф16mm costs 0.1 million yuan; (3) Replacing Ф16mm high-pressure hoses costs 0.2 million yuan; (4) Replacing three D80 cylinders equipped with buffers costs 0.54 million yuan; (5) Using 7 new electromagnetic valves with an effective diameter of Ф16mm and high flow capacity costs 0.35 million yuan ； (6) Replace 6 oil pressure cylinder inlet and return ball valves (Q21H—100DN20) at 0.3 million yuan each. (7) Add 1 external vertical shell-and-tube heat exchanger with a heat exchange area of 10 m2 at 30,000 yuan. The total investment amounted to 49,900 yuan; the system was put into operation for testing on February 6, 2005, and achieved the expected results. The operating times of the 14# gas stove’s oil pressure system before and after the modification are shown in Table 1, based on actual measurements of the relevant parameters. file:///E:/DOCUME~1/zzjjaa/LOCALS~1/Temp/ksohtml/wps_clip_image4.jpg For furnaces whose hydraulic systems have not been modified, the valve switching time remains roughly the same as that shown in Table 1 before modification; therefore, the measurement data before and after modifying furnace No. 14 are compared. As can be seen from Table 1, the direction-changing speed of each valve improved significantly after the modification. For the three-way valve, this time increased to 4.1 s, which is a reduction of 2.1 s compared to the 6.2 s before the modification ； The drop time of 2.7s is a 3.3s improvement over the 6s before the modification. After the modification of the chimney valve, the startup time is 4 s, which is an improvement of 1.6 s compared to the previous 5.6 s ； It takes 4.4s, which is a improvement of 7.4s compared to 11.8s before the modification. After the modification of the blow-off valve, the startup time is 3.0 s, which is an improvement of 1.5 s compared to the previous 4.5 s ； It takes 2.4s, which is 1.6s faster than the 4s before the modification. The direction-changing speed of the remaining Dg300 plate valves also increased significantly; due to the short stroke, the time required for them to reach their target position was very brief, so their operating parameters could not be measured accurately. In mid-March 2005, taking advantage of the annual maintenance period, one external vertical multi-tube heat exchanger with a heat exchange area of 10 m2 was installed on each return oil pipeline in oil pressure pump stations 4# to 6# of the new gas generation system. After it was put into operation, the cooling effect was quite significant, with the oil cooling temperature difference reaching 10–12°C. Obviously, the faster the operating speed of process valves, the more beneficial it is for production. Taking the example of the chimney valve of Furnace No. 14, whose operation time increased by 1.6 seconds after the modification, and assuming that the gas furnace operates normally for 330 days per year with 45 minutes allocated each day for shutdowns, ash removal, and cleaning, Furnace No. 14 can gain approximately 68 additional hours of effective gas production time per year. Based on an average annual gas production volume of 6,200 m3/h for this furnace in 2005, this translates to an increase in gas production of around 420,000 m3 per year. At a rate of 2,000 yuan per ton of synthetic ammonia produced from 3,300 m3 of gas, this equates to a value of 255,000 yuan in RMB. After the modification of the hydraulic system, the direction-changing time of the process valves was reduced by at least 1.6 seconds, achieving the desired outcome. Although the investment in optimization and modification is around 50,000 yuan, the heat exchangers, integrated blocks, connectors, etc. represent one-time costs, and they will not need to be replaced for a long time. After the modification of a single furnace, it can generate an additional value of 255,000 yuan per year; the entire investment cost can be recovered within just one quarter after the modification. Therefore, this modification is successful and economically viable. Through careful planning in the workshop, the optimization and modification of the oil pressure systems for the 16 gas generators in the gas production workshop were completed by November 2005, achieving the desired results. 5 Problems Existing After the Modification and Solutions: After the optimization and modification of the hydraulic system, two problems arose: (1) The direction-changing speed of each process valve increased significantly, which in turn led to a greater impact force on these valves. Although we used hydraulic cylinders with buffers to reduce the impact on the valves, this issue persisted. However, the buffering effect was poor. After repeated inspections on site, it was found that the height of the cylinder bracket was insufficient, which affected the performance of the hydraulic cylinder with buffering function. By adjusting the height of the cylinder bracket, a significantly better buffering effect was achieved. (2) When the oil pressure pump station is operating while the associated gas furnaces are shut down, the oil temperature in the pump station tends to rise slowly. Through research, it was found that the issue arises because the solenoid valve does not change direction when the gas stove is shut down. As a result, the oil cooler located in the return oil pipeline cannot provide cooling. Since the inlet and outlet pipes of the cooler have large diameters and low resistance, a large volume of oil flows through them; whereas the internal coiled tubes of the cooler have small diameters and high resistance, resulting in a smaller flow rate of oil. This is why the oil temperature in the pump station rises slowly. The workshop took prompt corrective action: a DN32mm ball valve was installed on the inlet pipe of the oil cooler at the pump station, and the workshop operators were instructed to close this ball valve whenever the gas furnace was shut down. The issue of the gradual increase in oil temperature at the pump station was completely eliminated through corrective actions.