[align=center]The Reasonable Application of Converter Pump and General Pump in Parallel Bus Systems[/align] Abstract : Based on the principle of pump speed adjustment and combined with the actual application in power plants, the feasibility of parallel operation of converter pump and general pump is discussed. Keywords : converter-pump and general-pump parallel bus systems 1. Introduction In water systems with a common bus system, the low efficiency of mixed use of converter pump and general-pump has always been a dilemma between saving electricity and saving money. In almost all bus system systems, the energy-saving effect of converter pump is calculated at the expense of efficiency or by increasing the burden on parallel general-pump pumps. However, in a power plant in Jiangsu, a reasonable method for the operation of general-pump and converter pump in a common bus system was found by combining previous energy-saving adjustment measures. Taking the circulating water of the turbine condenser in this power plant as an example, the relationship between the economical operation mode of the turbine and the flow rate of the circulating water pump is shown in Figure 1. The circulating water pump operates continuously with the unit for a long period of time. Due to the seasonal changes in the Yangtze River water temperature and the changes in the unit load, even under the same load, different external environments result in different circulating water flow requirements. Therefore, it is necessary to adjust the circulating water flow rate in a timely manner to ensure the safe and economical operation of the unit. The turbine vacuum is mainly controlled by adjusting the cooling water flow rate. To improve the economic efficiency of the unit operation, the increase in turbine power ΔN1 due to the increase in vacuum should be greater than the additional power ΔN2 consumed by increasing the circulating water flow. Net power increase ΔN = ΔN1 - ΔN2 ΔN1 is the turbine power increase; ΔN2 is the additional power consumed by increasing the pump flow rate; Dw is the cooling water flow rate; p is the condenser vacuum of the turbine. ΔN increases with the increase of cooling water flow. When it reaches point a, if the cooling water flow rate is further increased, ΔN begins to decrease until it reaches zero. When point c is reached, the expansion capacity of the turbine's final-stage nozzle has reached its limit, and the turbine power will not increase further; therefore, point c is the ultimate vacuum. The vacuum at point b, corresponding to the intersection of the isobaric flow line drawn from point a and the condenser pressure line, is the most favorable vacuum (Peco). The cooling water flow rate (Deco) at point a is the optimal cooling water flow rate. This demonstrates that properly adjusting the circulating water flow rate is a crucial means of improving turbine efficiency. 2. Applications in applications where pump blade angles can be changed Generally, the system's service pressure does not change significantly. Changes in pump house pressure are mainly determined by the water supply flow rate and the pipeline friction coefficient. However, the pipeline length, diameter, and material vary among different types of pump houses, resulting in different pipeline characteristics. The same pump operating at different speeds in different systems will have different operating points and significantly different efficiency. If the system's flow rate and head relationship follows a similar law to that of the pump, then within the allowable adjustment range, the pump's efficiency will remain at the same point. If the pump is properly selected, it can maintain maximum efficiency throughout the entire speed range, achieving the best energy-saving effect during pump speed regulation. If the relationship between the system's flow rate and head does not follow a similar pattern to that of the pump, then within the allowable adjustment range, the pump's efficiency will change with the speed, narrowing the pump's high-efficiency speed regulation range and reducing its energy-saving effect. When the pipeline characteristic curve is a horizontal line (i.e., there is no pipeline at the pump outlet), the pump can only adjust its speed within a very small range. Therefore, changing the pipeline characteristics and rationally configuring the pump characteristics are key to the efficient operation of a common main pipe network. In a main pipe water supply network, the currently widely used constant pressure control system, which controls the pump house outlet pressure, will cause the pump to operate under constant pressure when used on the distribution pumps. The pump's operating condition is equivalent to operating in a system with a high geometric head. The pump's efficiency changes along the Q-η curve with the flow rate, failing to expand the high-efficiency range, resulting in a very narrow high-efficiency speed regulation range. The overall efficiency of the pump is low, failing to achieve energy saving and consumption reduction, and only playing a role in flow regulation. Because the pressure it controls cannot directly reflect the pressure of the pipeline network, its effect on stabilizing the network pressure is not significant. A power plant in Jiangsu Province has two 135MW generating units. The circulating water is supplied by a single main pipe from the #3 pump house on the Yangtze River, using three 1000kW circulating water pumps. This pump combination was designed for two 125MW generating units. The normal design configuration is two operating and one standby. Traditional pump station designs focus solely on maximizing water supply, using different pump combinations of varying specifications and capacities to provide different flow rates and heads. In this case, a decrease in flow rate leads to an increase in head, resulting in energy waste. In 2002, the plant underwent a low-pressure cylinder expansion and renovation, increasing the unit capacity to 2×135MW. The cooling water volume increased accordingly. To adapt to seasonal changes and load fluctuations, improve turbine operating efficiency, and meet the different temperature and flow requirements of the turbine condenser, the plant adjusted the blade angle of its three pumps from +2° to -8°. This achieves both regulation and energy saving. Figure 2 shows the H-Q curve of the pump blade angle from -8° to +2°. In the diagram, n1 represents the pump blade angle +2° H-Q curve, n2 represents the pump blade angle 0° H-Q curve, and n8 represents the pump blade angle -8° H-Q curve. After modification, the pipeline resistance characteristics remain unchanged, but the water flow rate decreases and the water pressure drops at the same speed. During summer (July to September), two industrial frequency pumps are combined as n1 and n2. The diagram shows that when the two pumps operate at 0° and +2° respectively, the water output Q = Q1 + Q2, with the +2° pump often operating under overload. The motor currents are 87A for the pump with a blade angle of 0° and 130A for the pump with a blade angle of +2°. When both pumps operate at +2°, the water output is 2Qn, the outlet water pressure is 0.125 MPa, and the motor current is 124A. In both operating modes, 2Qn ≈ Q1 + Q2, and the total operating current 87 + 130 < 2 × 124. Therefore, with a constant input voltage, the power consumption is 2 × (+2°) > 0° + (+2°). When both pumps are operating at a blade angle of 0°, the water output is approximately 2 × Q1 < 2Qn, and the operating current is 104A. This method is insufficient for peak summer loads, especially after the 125MW unit was upgraded to 135MW. It is frequently necessary to operate the standby pump to supplement the insufficient water volume. Therefore, to save energy and ensure sufficient system backup capacity, the commonly used pump combinations are n1 and n2 (0° + (+2°)), or 2 × (+2°) combinations. In winter, during the late-night low-peak load, operating one (+2°) pump is sufficient. During peak loads, the maximum operating flow rate of a single pump is insufficient; before the old unit was upgraded, a 300kW pump from the old unit was used as a supplement. The total current I = 110A + 26A = 136A. Since the old unit was dismantled in 2002, it has been necessary to operate two 1000kW pumps. The operating mode is as follows: both pumps are adjusted to their minimum blade angle (-8°) (operating point Q3), the water volume meets the peak load requirements, and the motor operating current is I = 2 × 70A = 140A. This is close to the working condition of a large pump plus a small pump, which is relatively ideal. The power factor at this time is only 0.63. However, during the evening off-peak load, it is necessary to start the pump (+2°), stop both pumps (-8°), and then adjust the blade angle of one of the pumps to +2° as an emergency hot standby, which is relatively cumbersome. In order to meet the needs of energy saving and facilitate adjustment according to load, a Diamond-HV06/1250 high-voltage frequency converter was installed on pump #9. The Diamond-HV frequency converter has the characteristics of wide adjustment range, strong overload capacity, and low power consumption. It is particularly suitable for on-site installation in the machine room where a separate equipment room cannot be built. The pump speed regulation shown in Figure 2, obtained through operational testing, is based on the pump similarity principle (n/n1=Q/Q1, n2/n12=H/H1, n3/n13=P/P1, where n is the pump's rated speed, n1 is the pump's adjusted speed, H is the pump's design head, H1 is the pump's adjusted head, Q is the pump's design flow rate, Q1 is the pump's adjusted flow rate, P is the pump's design shaft power, and P1 is the pump's adjusted shaft power). By changing the pump's speed, its characteristic curve changes, and the pump's operating point also changes accordingly. This allows the pump's characteristic curve to intersect with the system's pipeline characteristic curve at the required flow rate, avoiding excess head caused by valve throttling or increased pipeline pressure. Simultaneously, it may bring the pump's operating point, which has exceeded its high-efficiency zone, back into the high-efficiency zone. Pump speed regulation can only change the pump's characteristic curve, not its performance or pipeline characteristics. The energy-saving effect of pump speed regulation in engineering lies in reducing the excess head of the pump or changing its operating point, bringing it back from the high-efficiency zone. Therefore, the feasibility of modifying the pump's characteristic curve within the pipeline network to adapt to changes in service point parameters is primarily determined by the desired pump head. The plant's three circulating water pumps, designed with a head of 18 meters, allow the pipeline water pressure to fluctuate between 0.085 MPa and 0.12 MPa, thus providing ample room for energy saving through pump configuration and speed regulation. In the figure, n1, n2, and n3 represent the H-Q curves of the variable frequency pump at different speeds, where n1 > n2, and n3 > n3. The N curve represents the H-Q curve of the fixed frequency pump with a changed impeller angle (-4°). The blade angle of pump n1 is +2°. The H-Q curve of the variable frequency motor operating at 50Hz is consistent with the curve of the mains frequency motor operating at 50Hz. During summer, under normal operation, at peak times, the variable frequency pump operates at Qn, and the mains frequency pump operates at Q1. During off-peak times, both the variable frequency and mains frequency pumps operate at Q2, with their outputs being roughly equal. At this time, the variable frequency pump operates at 42Hz, both within their high-efficiency ranges. The turbine condenser inlet pressure can fluctuate within a certain range, and the heat dissipated is related to the turbine exhaust volume, cooling water temperature, and flow rate—that is, the water flow rate under the same season and load. While meeting the mains water pressure, when the frequency is reduced below 40Hz, the mains frequency pump's output is greater than the variable frequency pump's, and its efficiency begins to decrease. In this system, the mains frequency pump has the minimum blade angle, and the variable frequency pump has the maximum blade angle. The variable frequency pump's operating frequency can be adjusted down to a minimum of 35Hz. Further reducing the frequency will result in the variable frequency pump failing to output water due to excessive pressure difference and will cause vibration due to pump stalling. The optimal operating range for the two pumps in this configuration is 42Hz to 48Hz. To accommodate the changing circulating water volume requirements during winter and to achieve an ideal configuration of the fixed-frequency and variable-frequency pumps, the impeller angle of the fixed-frequency pump is generally -6° (H-Q curve N'), while the variable-frequency pump remains at +2°. This combination shifts the combined efficiency curve to the left, ensuring operation in the high-efficiency range even at lower frequencies. Furthermore, it prevents pump stalling above 30Hz. During low-peak loads, the variable-frequency pump operates independently, adjustable within the 42Hz to 50Hz range. If the generator set operates alone, it can operate within the 38Hz to 46Hz range. During peak loads, the two pumps (-6° + (+2°)) are combined, with the variable-frequency pump operating at 45Hz and a current of I = 68 + 70 = 138A. Thus, the variable-frequency pump operates as the main pump, with one at +2° and the other at -6° as backup. This satisfies the operating requirements while significantly improving adjustment methods, energy-saving potential, and circulating water backup capacity. 3. Application in Pumps with Non-Adjustable Blade Angles For pumps with non-adjustable blade angles, if the pump is connected to a main pipe via a control valve, the relationship between the main pipe pressure and the pump flow rate changes. Figure 4 illustrates this relationship. The original flat characteristic n1 becomes n2 after pressure reduction via the valve, indicating that the pump's pressure characteristic n2 on the main pipe becomes steeper. If the outlet valve of the variable frequency pump is fully open, the pump still exhibits characteristic n1. When the variable frequency pump operates at adjustable speed, the H-Q characteristic curve shifts downwards parallel to n3. Thus, by partially closing the valve to control the steepening of the characteristic curve, both fixed-frequency and variable-frequency pumps can find a pressure equilibrium point on the main pipe over a wide range, increasing the speed range of the variable frequency pump. However, this results in greater energy loss at the valve. In Figure 4, the product of the pressure loss between the n1 and n2 characteristics and the corresponding flow rate represents the energy loss at the valve. [1] As can be seen from the above, when a variable frequency pump and a fixed frequency pump with the same characteristics are connected to the grid, it is difficult for the variable frequency pump to reduce the frequency for water supply when the original characteristics are relatively flat. For example, if the fixed frequency pump has already been artificially made to have a steeper valve characteristic before being connected to the grid, although it can reduce the frequency and reduce the water supply, the network pipe cannot be stabilized and the energy saved by the variable frequency pump speed regulation may not be able to compensate for the increased losses of the valve. The solution is to increase the pressure and capacity of the variable frequency pump to increase the adjustable water supply, improve efficiency and maintain constant pressure in the network. Maintaining constant pressure in the network is not a completely necessary method to effectively regulate the pump to work in the high-efficiency range. Although it can make each pump connected to the grid work at a given pressure, some pumps may produce very little water or no water at all. Especially for axial flow pumps, vibration will occur at the critical point of pump stall, which seriously affects the stability of the network and the safety of pump operation. On the other hand, pay attention to whether the motor load rate is appropriate. If the speed is low, even if the pump efficiency is high, if the motor load rate is too low, the overall efficiency will be very low and the purpose of energy saving will not be achieved. The ideal control method for pump speed regulation is to ensure that the pump's water supply can maintain the service pressure of the pipeline network, and that the change in pump speed makes the pump's QH curve match the pipeline characteristic curve. In this way, the pump will always work in the high-efficiency zone within the effective regulation range. In reality, it is difficult to achieve such regulation. Therefore, a control method that simulates pipeline characteristics can be adopted, that is, the pipeline characteristic curve measured in practice is input into the control software, and the pump speed is automatically controlled by two variables: pump outlet pressure and pump flow rate, so that the pressure and flow rate change according to the input curve. [2] 4. Conclusion From the above two situations, it can be seen that for systems with steep H-Q characteristics or where measures are taken to change the pump's H-Q characteristics in the pipeline network, that is, where the pipeline pressure loss accounts for most of the total head or the service pressure changes significantly, pump speed regulation can be adopted. After frequency conversion speed regulation control, the pump can maintain or reduce the outlet head at low speed and low flow rate, thus achieving high-efficiency and energy-saving operation. For systems where pipeline water pressure loss accounts for a small proportion of the total head and the service pressure changes are small, pump speed regulation may not be economically reasonable; for systems where pipeline water pressure loss accounts for a very small proportion of the total head and the service pressure changes are very small, pump speed regulation is not recommended.