WO2005054680A1 - Centrifugal pump - Google Patents

Abstract

A centrifugal pump consists of an impeller, a fluid-guiding mean, a pump casing and shaft parts. Fluid channels are formed between the impeller vanes, wherein the end section of each fluid channel is curved in the sense of opposite tangent direction of rotation and its cross section area is reduced gradually. Fluid is accelerated and changed direction at the end of the channels where the centrifugal force effected. Then the fluid with high relative speed and outlet angle which near 0 degree is discharged from the impeller. An absolute speed at channel outlet is reduced. A moment of counter force produced by changing direction and acceleration depresses the power of the rotating shaft. The above structure remarkably improved the ratio of kinetic energy and potential energy of the pump. The configuration at the end of the impeller equals to continuous opening, but no circumfluence interference absolutely. Cavitation resistance and high efficiency are improved respectively.

Description

 High potential ratio, internal friction reduction, centrifugal booster centrifugal pump

 Method and example for technology combination thereof

 The present invention relates to the improvement of the structure and principle of a centrifugal pump, as well as methods and examples of technical combination thereof. Improve the design to form a high potential ratio, internal friction reduction, centrifugal booster centrifugal pump product that reflects high performance in terms of throttling, efficiency, power, and cavitation, or at the same time in terms of volume, cost and ease of use Advantage. The new method is used for modular combination between new technologies or between new technologies and existing technologies, which can form more types of new centrifugal pump products.

Background technique

 The centrifugal pump consists of an impeller, a deflector, a casing, and a shaft system component. Among them, the impeller is a component that drives the liquid flow to perform work, and the deflector is a component that converts liquid flow energy into pressure potential energy. Centrifugal pumps are the most widely used fluid machinery and are widely equipped in various industries of the national economy. In the fields of mining, electric power, petroleum, chemical industry, machinery, light textile, environment, urban and rural construction, and water conservancy, in the traditional agriculture, forestry, animal husbandry, and fishery industries, centrifugal pumps are all commonly used equipment, and their installed capacity is very huge. For example, the installed capacity of agricultural irrigation and drainage pumps in China alone exceeds 100 million kilowatts, most of which are centrifugal pumps.

 Centrifugal pumps were invented in France at the end of the 17th century and have a history of more than 300 years. Its basic equations were derived in the middle of the 18th century and are still the common basis for the evolution of new theories. The history of water pumps being used by mankind as the largest public driving device is accompanied by industrial civilization and urbanization. Some people even compare it to the heart of the national economy, but most people, including the inventors, do not know . The water pump mentioned here mainly refers to the centrifugal pump.

Since the 19th century, especially at the end of the 20th century, the research on centrifugal pumps has been hot. Its design and production process have undergone several important improvements, and its performance has been continuously improved. However, the existing technology products are still not satisfactory, and their efficiency is still low. The cost is still high, and its installation and adjustment mode often cannot meet the needs of users. The main reasons for the low efficiency of centrifugal pumps are the following:

 First, the impeller output fluid velocity is too high and the kinetic energy ratio is too large. The hydraulic loss of kinetic energy in the deflector is proportional to the square or even the cube of the speed. When the speed is too high, most of the kinetic energy will be lost.

 Too high a flow rate causes a reduction in the efficiency of the centrifugal pump, which is well known in the academic community. All academic books recommend the backward curved impeller in order to reduce the speed. Most of today's centrifugal pump products use this impeller, but the problem remains unsolved.

It can be seen from the following analysis that the backward curved blade does not reduce the flow velocity much. According to the exit velocity triangle, the absolute velocity v ₂ of the impeller exit flow is the vector sum of the exit implication velocity U ₂ and the relative velocity W _2. Since W ₂ is an order of magnitude smaller than U ₂ , the effect on V ₂ is small. In the commonly followed velocity diagrams, the graphic proportion of the relative velocity ^ is often exaggerated (the picture is too small to see clearly), so that it seems that a backward-curved blade with a smaller exit angle β ₂ can reduce the absolute velocity V ₂ much. In fact, due to the large area of the impeller outlet, the outlet relative velocity W ₂ than the peripheral velocity U ₂ is small. A triangle constructed in true proportions shows that the adjacent edge u _{2 of the} exit angle p ₂ is large, while the other adjacent edge

, Change ₂ does not significantly change the length of its opposite side v ₂ . Therefore, the Euler equation Y _T = u ₂ v ₂ cosa ₂ -uivicosax = 0.5 (u ₂ ² -ui ² + wi ² -w ₂ ² ) + 0.5 (v ₂ ² -V! ² ) The specific potential energy term of 0.5u ₂ ² and the specific kinetic energy term of 0.5v ₂ ² are very close in value. Since the values of the other specific energy terms are small, the impeller output potential-to-momentum ratio is actually close to 1.

 The high-speed diversion increases the diversion load and reduces the hydraulic efficiency of the diversion process. This is the primary reason for the low hydraulic and overall efficiency of the centrifugal pump.

Second, in the impeller process, the velocity distribution of the flow field is uneven and unstable, and turbulence, backflow, and outflow have always been difficult to eliminate, which seriously affected the hydraulic efficiency of the process. At the junction of the impeller process and the diversion process, the liquid flow is in a severely under-constrained state, so that the relative vortex abduction, turbulence, backflow, and off-flow driving force mutual feed, The area of influence is enlarged, and local incentives are severe. All these factors lead to increased losses.

 Thirty-three, the friction loss of the impeller disk occupies a proportion that cannot be ignored, which reduces the internal mechanical efficiency. This loss has nothing to do with the flow velocity, but it is strongly related to the impeller diameter and its high power. As the impeller diameter or speed increases, the losses increase dramatically. For centrifugal pumps with low specific speeds and high heads, disc friction may cause a reduction in efficiency of more than 10%.

 The above three major reasons, combined with the existence of other types of losses, determine that the centrifugal pump may not have a satisfactory efficiency index. According to the data of the inventor with more than 10,000 models and specifications (the product sample of a Chinese company is surprisingly and thoughtfully over 5,000), the nominal efficiency of the prior art products is mostly around 50%, the highest The nominal value is 82%, the lowest value is only over 30%, and the average value is between 50% and 60%.

 Nominal efficiency is an index that can only be achieved when operating under optimal conditions. In fact, due to the difference between the discrete distribution of the pump's specifications and the continuous distribution of actual demand, and due to possible changes in the application conditions of the pump, the main parameters such as pressure and flow often need to be adjusted by external means during operation, which makes It is very common for the centrifugal pump to deviate from the design conditions, which causes the statistical distribution of the actual operating efficiency of the centrifugal pump to decrease. Its roots are:

 Fourth, the structural design of the impeller and the deflector did not take into account the widely existing practical needs of variable operating conditions. A flow channel design that is sensitive to the flow direction and has a loss characteristic optimized for the design flow is generally used. For example, The inlet and outlet angles of the impeller blades, the angles of the guide vanes and the counter guide vanes of the impeller, etc. In the actual variable operating conditions, the change of the flow rate causes a change in the direction of the speed, and the change in direction produces impact losses, resulting in the actual efficiency being lower than the optimal operating efficiency.

The use of external resistive adjustment, which deviates from the design conditions, not only causes an increase in the internal loss of the pump, but also an external energy loss. Users actually It has withstood the dual energy losses that occurred internally and externally. External losses cannot be expressed in the pump efficiency index, and there is no separate measurement display and statistics. In order to calculate this loss, the unit efficiency and the energy utilization rate of the entire flow system should be examined, which are obviously lower than the pump efficiency.

 Water pumps are among the top energy-consuming equipment in modern society, and their inefficient operation has caused huge socio-economic losses. In China alone, the annual power consumption of water pumps exceeds 400 billion kWh, and the part that can be tapped to save potential is conservatively estimated to be over 50 billion kWh. This increases the operating costs of the national economy and also increases the environmental pollution load in energy production. The efficiency of centrifugal pumps is a major technical and economic issue that must be addressed, and its urgency is widespread worldwide.

 In addition to the urgent need to improve efficiency, reducing size, tube structure and process, reducing manufacturing costs, increasing functional value and other issues related to cost performance are also generally concerned by the pump industry and users from all walks of life.

 The task of the present invention is to overcome the above-mentioned disadvantages of centrifugal pumps and further create new value.

Summary of the invention

 The first specific objective of the present invention is to: optimize the specific energy attribute structure of the impeller output, design a new impeller to realize such optimization and improve its flow field characteristics, so that the hydraulic efficiency of the guide flow and the impeller stroke is greatly improved at the same time.

 The second specific object of the present invention is to change the friction medium on the outside of the impeller wheel disc of the centrifugal pump and reduce its viscosity coefficient, thereby greatly reducing the friction loss of the disc wheel and improving the internal mechanical efficiency of the pump.

 The third specific object of the present invention is to improve the structure of the deflector and other components to match the overall design, so as to further reduce the diversion loss and improve the overall hydraulic efficiency, and to reduce the deflector fundamentally. Compared with the case size, the manufacturing cost is lower and the use is more convenient.

The fourth specific object of the present invention is: to design a method for Among the components, the most effective technical combination is made between the new-type components and the traditional technical components to produce positive effects in terms of efficiency, cost, or use function. Since the present invention is premised on the improvement of related technical theories and the innovation of technical concepts, the necessary arguments in this regard are also attached to the purpose of each invention.

 The technical route for achieving the first objective of the present invention is: design a new type of centrifugal pump impeller, the output of the impeller has a higher potential head and a lower dynamic head, and the ratio of the two is measured as the potential-dynamic ratio. Parameters, so that the ratio is significantly improved compared to the prior art. Increasing the potential-to-motivation ratio reflects the purposeful design of transforming the impeller into a component that mainly converts the shaft work into pressure potential energy, because there is no relative speed factor in the line integral supercharging equation relying on the centrifugal force and the path dot product in the impeller flow path. Producing as much potential energy as possible in relatively low speed impellers can reduce the rate of flow loss. The further reduction of dynamic head means that the absolute speed of the impeller output must be further reduced based on the limit value of the prior art backward curved blade speed reduction scheme. The advantage is that the hydraulic loss of subsequent deflectors decreases with the decrease of flow velocity. Rapid reduction, reduction of kinetic energy ratio and reduction of kinetic energy loss rate will result in a linear increase in pump efficiency.

 The momentum ratio has a monotonically increasing correspondence function with the reaction degree or reaction coefficient in the prior art concept. Therefore, increasing the momentum ratio is to increase the reaction degree or reaction coefficient. The concept of momentum ratio is defined and used in the present invention because it is easier to understand, and can more clearly and specifically express the characteristics and value of the present invention.

 As mentioned earlier, there is a speed limit for backward curved blades. To overcome this limit, prior art impellers must be retrofitted. These transformations will include new designs such as changing large outlets to small outlets, and further reducing the angle of the liquid flow outlet to equal to or almost equal to zero, in order to produce relative speeds of outlets with favorable directions and sizes. The purpose and significance of these transformations is to make proper weighting arrangements for two potential energy production pathways with different loss rates so that the efficiency of the centrifugal pump can obtain an optimal non-linear combination of expressions.

The technical solution of the present invention to achieve the first object of the invention is: adopting a high-potential-dynamic-ratio impeller, the tail of the impeller channel of the impeller is curved in the tangential direction and the cross-sectional area is gradually reduced, The fluid is accelerated and changed direction at the end of the path of work performed by centrifugal force, and finally flows out of the impeller at a relatively large speed and an exit angle close to 0, the absolute speed of the outlet is correspondingly reduced, and the reaction torque generated by the steering and acceleration processes reduces the work of the rotating shaft.

 The solution of the invention is designed on the basis of the law of conservation of mechanical energy. The energy conversion at the end of the leaf groove process in the rotating coordinate system to change the output potential-to-dynamic ratio is a strategic application that favors the potential energy production environment and avoids disadvantages. The applied dynamics is resistance, and the conversion process is low. Lossy.

The principle of the present invention is deduced from the Euler equation describing the specific work of the impeller, which is the result of optimizing the system term of the equation. Consider the equation Y _T = u ₂ v ₂ cosa ₂ -UiViCosai = 0.5 (u ₂ ² -u! ² ) + 0.5 (W! ² -w ₂ ² ) + 0.5 (v ₂ ² -vi ² ), which describes The quantitative relationship between the impeller's energization of the liquid flow motion composed of the implicated motion in the stationary coordinate system and the relative motion in the rotating coordinate system. Among them, 0.5 (u ₂ ² -ui ² ) is the line integral value of centrifugal force work converted into specific potential energy, which is completed on the absolute motion path of the centrifugal force field in the stationary coordinate system. 0.5 (wj ² -w ₂ ² ) + 0.5 (v ₂ ² -vi ² ) Contains all relevant terms of the specific kinetic energy of the motion synthesis in the two coordinate systems, implicit mechanism of kinetic energy and potential energy exchange, and observance of the law of conservation of mechanical energy (Excluding friction loss). The specific kinetic energy increase of the liquid flow in the impeller flow channel is 0.5 (v-v), which is directly derived from the normal force work of the blade (equal to the remaining work after deducting the supplemental centrifugal force work energy), and the specific potential energy increase is 0.5 (u ₂ ² -u! ² + wi ² -w ₂ ² ) are derived from the centrifugal force work and the specific kinetic energy reduction of relative motion, respectively.

For the purpose of increasing production capacity, the prior art has always set the relative motion specific kinetic energy reduction term O iw-w ₂ ² ) to a positive value, because doing so can cause a simultaneous increase in specific work and potential head. However, traditional ideas place too much hope on the exit velocity triangle synthesized by motion. In fact, the law of cosines formula ^{_{v 2 2 = u 2 2 +}} w 2 2 - 2u 2 w 2 cosp 2 can not give the desired junction, since w ₂ u ₂ than an order of magnitude, such that w _{² 2} - 2u ₂ w ₂ cosp ₂ is almost equal to 0 with respect to u ₂ ² , so there is _V2 u ₂ ² . Let the absolute high speed of the Fgen \ ^ ₂ enter the diversion process, which will produce huge Large losses, so that the first major bottleneck restricting pump efficiency. The only way to solve this problem is to change the relative motion ratio kinetic energy decrement term 0.5 (W! ² -w ₂ ² ) to a negative value, that is, to increase w ₂ through inverse energy conversion. At this time, the above cosine theorem formula will be Given a lower absolute speed v ₂ , the efficiency bottleneck of the centrifugal pump can be broken.

 Based on the above analysis, the main design feature of the present invention—an exit angle close to zero and a relatively large exit relative speed—are justified.

When the exit angle β ₂ 0, the exit relative speed w _{2 is} opposite to the implication speed u ₂ , which can offset u _{2 to the} greatest extent. Setting β ₂ «0 has another important effect, that is, in the case where the outlet area is reduced and the outlets are separated from each other, a small outlet angle is required to organize the velocity field after the liquid flow exit to avoid disadvantages. The velocity distribution causes turbulence. u ₂ give _{2 = (1- K) - β} Ο means exit velocity triangle reduced to straight line segment, an absolute velocity by algebra ¥ ₂ = 11 _2. In the formula, the coefficient Kw ₂ / u _{2 is} called a feedback reduction ratio. K is an important control parameter, and its size linearly reflects the degree of absolute speed reduction. The reason why it is called the feedback reduction ratio is that when the inverse tangential relative speed w ₂ is generated in the rotating coordinate system, the reaction torque reduces the work of the rotating shaft, which is a lossless kinetic energy feedback process. The kinetic energy feedback deceleration appears as acceleration in the rotating coordinate system, so there must be an acceleration flow path in the impeller structure.

In order to significantly reduce the absolute speed v ₂ , W ₂ must be set to be much larger than the limit value of the prior art in value, and should reach a level of the same order of magnitude as u ₂ . This can be achieved by reducing the exit area and setting an accelerating section that gradually reduces the cross-sectional area at the end of the blade channel. When K = about 0.5, w ₂ = Ku ₂ 0.5o) R ₂ . The increase in the relative speed w ₂ is necessarily at the expense of potential energy, so the output specific potential energy increase of the impeller will decrease as w ₂ increases. Since the decrease of the specific potential energy increment is 0.5o ² R ₂ ² K ² , which is proportional to the square of the true decimal K, the value is small. The reduction in specific kinetic energy increase is equal to 0.55 ₂ ² -0 · 5 (u ₂ -Ku ₂ ) ² = 0.5ω ^{2 2} (2Κ -Κ ² ), which is much larger than the reduction in specific energy. The ratio of the two is (2-K) / K, so the output momentum ratio of the impeller will increase significantly. Calculations show that the practical interval The value of K can increase the output momentum ratio by 2 to 8 times.

 The solution of the invention has a significant effect on improving the efficiency of the pump. The output potential of the high potential ratio impeller is relatively small, and the diversion load is light. The speed power loss will be greatly reduced, so it has a much higher diversion efficiency than the prior art.

 Most of the hydraulic loss of centrifugal pumps occurs in the deflector. The inlet flow rate of the prior art centrifugal pump is very high, usually exceeding 20 m / s, which is an order of magnitude higher than the hydraulic specification, which inevitably results in large losses. In deflectors, high-speed marginal friction and large-gradient internal friction are unavoidable; high-speed impacts, deflows, and local excitations also often occur. These losses have a sharply correlated sensitivity to the power of 2 or 3 due to the inductive velocity. Diversion loss is particularly serious in small and medium-sized, low specific revolutions, and deviations from design conditions. At this time, the kinetic energy loss rate may exceed 50%. The present invention has a particularly significant effect on rescuing these losses. The following is a detailed discussion of the effects of this effect by distinguishing between two types of diversion losses, the local resistance type and the along-path resistance type.

 For local resistance-type diversion losses, such as turbulent flow regimes such as impacts, local excitations, and sudden changes in velocity, the losses are concentrated locally. This type of loss is proportional to the square of the flow velocity and proportional to the local drag coefficient. For tube analysis and direct comparison, the ratio of this type of conduction loss of high potential ratio liquid flow to constant potential ratio liquid flow (potential ratio is 1) is defined as the local resistance type conduction loss ratio, and this ratio depends on the former's control parameters. K changes are listed in the third row of Table 1.

For resistance-type diversion losses along the way, such as the typical booster flow path losses of a deflector, its analysis and calculation rely on path integrals, and there are many parameters involved in the calculation. The premise of the comparison is as follows: the high potential ratio and the constant potential ratio have the same specific energy of the inflow, the same flow rate, the same cross-sectional area expansion rate, the same friction coefficient, the inflow velocity ratio is known, the geometry is similar, and so on. For cartridge analysis, it is also assumed that the speed of the diversion outlet is the same, and the effect on the result can be ignored. Based on these prerequisites, theories related to fluid mechanics can be used to derive: The ratio loss ratio (ratio of specific energy loss per unit length) of the high-potential ratio and the constant-potential ratio diversion section along the way is directly proportional to the square of the velocity ratio. Listed in Table 1 No. 4 The length ratio of the diversion flow, which is determined by the diversion load ratio of the two, is directly proportional to the ratio of the diversion speed. The data is listed in the fifth row of Table 1. Then use a similar method to process the local loss to define the two-way resistance type guide loss ratio along the way, ignoring the high-order small amount caused by the exit speed. The integral can be obtained that the loss ratio is proportional to the third power of the two guide speed ratio The data are listed in the sixth row of Table 1.

 Relation loss between high potential ratio liquid flow and constant potential ratio liquid flow and velocity measure K of the former

 Refer to Table 1, where the first row is the independent variable, which is the control parameter of the high potential ratio impeller

——Feedback reduction ratio K, is the speed measurement of the numerator of each loss parameter ratio. The second line is the typical value of the high potential specific liquid flow potential dynamic ratio, which is the specific energy property measurement of the ratio molecule. The remaining variable 3, 4, 5, and 6 rows of data are the ratios of the comparison parameters, which respectively represent the local resistance type diversion loss ratio of the two liquid flows, the resistance type diversion ratio loss ratio along the way, the diversion flow length ratio, and along the way. Resistance type diversion loss ratio.

 The data in the first column is all 1, because both are constant potential specific flow and the parameters are the same. The remaining columns are all true decimals, indicating that the high potential specific flow has a lower conductivity loss index, and these ratios decrease with the increase of the high potential specific flow velocity measure K. Lines 3 and 6 directly give the ratios of the local resistance-type diversion loss and the along-path resistance-type diversion loss. Both ratios are reduced in sharp correlation with K.

The data in line 6 shows that within the value range of K, the drainage loss of high potential ratio liquid flow is one order of magnitude lower than that of normal potential liquid flow. The practical significance is: For the resistance-type diversion along the way in the dominant mode, the loss of the high potential ratio liquid flow is smaller than the normal potential liquid flow by one. Magnitude. In fact, because the constant potential ratio diversion technology often includes some local resistance type losses mixed in, especially the variable angle impact type secondary type losses caused by variable flow operation exist universally and often, making the ratio type practical The denominator is larger, there is no such factor in the high-potential diversion characteristics, and there is a complete set of low-loss diversion technology support, so the actual ratio will be smaller or even much smaller than the data in Table 1, and this is left to the text . The conclusion is that the high potential ratio liquid flow output by the high potential ratio impeller has the characteristic of reducing the conduction loss by orders of magnitude.

The positive effects of the invention can be exemplified. For example, suppose that the potential-to-kind ratio of a prior art centrifugal pump is 1, and the kinetic energy loss rate of its local resistance type diversion is 50%, and the specific energy loss rate is 25%, which is ideal in all other losses. In this state, the efficiency of the pump will be only 75%. If the high potential ratio impeller of the present invention is used instead, and the comparison parameters such as the flow rate and the specific energy are the same, but the potential ratio is increased by 3 times, that is, from 1 to 4, the normalized guide load ratio of the deflector is 1 / (1 + 4) = 20%, reduced to 40% of the corresponding load ratio of the comparison pump 1 / (1 + 1) = 50%, and its conduction speed reduced to (0.2 / 0.5) ° · ^{5 of the} corresponding speed of the comparison pump -63.25%, the amount of kinetic energy loss and the specific energy loss rate proportional to it are reduced to 40% of the comparison pump according to the speed square law, but the kinetic energy and specific energy loss rates of the two are equal. Under the same assumptions, the specific energy loss rate is reduced to 25% x 40% = 10%, and its efficiency will be increased to 90%.

The above example analysis uses the insertion loss analysis method. Without loss of generality, remove the assumption that "all other losses are 0,", the introduction of the concept of diversion insertion efficiency can easily analyze the law of diversion loss separately. Diversion insertion efficiency is the diversion hydraulic efficiency, which is defined as the diversion process. The ratio of the specific energy of the output to the specific energy of the input. In the efficiency equation, it is a non-linear factor of the overall hydraulic efficiency and total efficiency. Ignoring high-order small quantities can be regarded as a linear factor. In this analysis method, the momentum is It is more convenient to select the variation parameter, because the loss of specific kinetic energy is proportional to the square of the flow velocity, that is, proportional to the specific kinetic energy. Note that in the example, the loss rate of kinetic energy and specific kinetic energy does not change with the potential kinetic ratio. Introduce the proportionality constant, normalize the loss kinetic energy contrast energy, and get a high potential ratio liquid flow bureau The efficiency formula of partial resistance type diversion is shown in formula (1).

Where ^ is the specific kinetic energy loss coefficient of the local resistance type diversion at the design flow. ^ / (1 + λ) is the specific energy loss rate, which is directly proportional to the specific kinetic energy, and is therefore inversely proportional to (l + λ). As mentioned above, the potential ratio of the prior art centrifugal pump is basically 1, so its efficiency is ii _hinl = 1-i _{a. After} switching to a high potential ratio impeller, its efficiency increases according to the formula (1), which increases Efficiency performance is shown in Table 2.

Table 2.Comparison of local resistance type diversion efficiency of high potential ratio and constant potential ratio liquid flow

It can be seen from Table 2 that even for a relatively poor deflector, such as a deflector with a specific kinetic energy loss coefficient of ₁ = 0.5, when the potential-to-kind ratio λ = 3 to 9, its local resistance type diversion efficiency will be 75% of the technology increased to 87.5% ~ 95.0%. It can be seen that increasing the potential-to-dynamic ratio can greatly improve the efficiency of local resistance type diversion.

Closer to reality is the resistance-type diversion loss along the way. The data and conditions in the above example are still used to analyze and compare this kind of loss. In the example, the diversion load ratio of the high potential ratio to the normal potential ratio is 40%, and the inducting speed ratio is 63.25%. The two estimated ratios are unchanged. The conditions of similar flow geometry, cross-sectional area expansion rate and friction coefficient are the same. The length of the diversion flow of the specific liquid flow will be reduced to 63.25% of the contrast pump, and the amount of kinetic energy loss per unit length along the route will be reduced to 40% of the contrast pump according to the speed-squared law, and the amount of kinetic energy loss will be reduced to 0.6325 ³ = 25.3%. Obviously, the specific kinetic energy loss rates of the two are no longer equal, and this index is directly proportional to the conduction velocity. Under the same assumptions, the specific energy loss rate will be reduced to 25% x 25.3% = 6.3%, and its efficiency will be increased to 93.7%.

 Regarding the resistance loss along the way, the specific kinetic energy loss rate is proportional to the inductive velocity. Therefore, only the maximum value of this velocity, that is, the specific kinetic energy loss rate of the potential-free liquid flow, can be used as a constant common constant. This constant is introduced, and the loss contrast can be normalized. The flow efficiency formula of the liquid flow with high potential ratio and constant potential ratio can be obtained as formula (2).

hini = 1-ξ ₂ / (1 + λ) ^1,5 (2) where ξ ₂ is a potential-free liquid flow (λ = 0) with the same flow rate and the same specific energy. The structure is similar, the expansion rate is the same, and the friction coefficient The kinetic energy loss rate of resistance-type diversion along the way under the same setting conditions and the same exit speed is a measure of the quality level of the structure and process of such deflectors. When λ is changed, the specific loss coefficient and flow length of the matching deflector are not the same, and the specific kinetic energy loss rate is also different, but the specific energy loss rates are all subject to the same constant ξ ₂ and are comparable. In the formula, ξζ / Οί + λ) ¹ · ⁵ is the specific energy loss rate of the deflector, which is directly proportional to the third power of the flow velocity, and is therefore inversely proportional to (1 + λ) ¹ · ⁵ . The potential ratio of the prior art centrifugal pump is 1, and its diversion efficiency is

-i-^ z ^ ii sse After switching to a high potential ratio impeller, the diversion efficiency improves according to the formula (2), and its synergistic performance is shown in Table 3. Comparison table of high-potential ratio and constant-potential ratio liquid flow resistance-type diversion efficiency along the way

Referring to Table 3, the increase of the efficiency data of the high potential ratio is significantly higher than that of Table 2. Even if it is a relatively poor or poor supporting deflector, such as a deflector with a reference loss coefficient ξ ₂ = 0.7071 ~ 0.9899, its normal potential is only 75% ~ 65% than the diversion efficiency of the diversion. When ≥4, the high potential ratio diversion efficiency is more than 91%. When ξ ₂ = 0.7071, the constant potential specific diversion efficiency is 75%, while the high potential specific diversion efficiency is as high as 93.7% to 97.8%. It can be seen that, for the resistance type deflector along the way, increasing the potential-to-dynamic ratio can significantly improve the diversion efficiency, and has an order of magnitude advantage. Called an order of magnitude advantage means increasing the l-axlO ^_n to ^1-axlO-n- 1 wherein a is a decimal, n is a positive integer.

Resistance-type diversion loss along the way is a regular loss mode of centrifugal pump deflectors, but local resistance-type loss is also unavoidable under the existing martial arts framework, especially the throttling adjustment in the actual fluid flow system will lead to variable working conditions Operation and local losses occur. Therefore, the discussion of the diversion efficiency of the present invention has to be elaborated, because their mechanical models are different, and the results are also very different. In order to obtain actual comparison data, the statistical total based on the statistical data of the two loss modes in the prior art deflector should be obtained. The weight coefficient of the volume is used to weight the two types of diversion loss and efficiency data, and the result will be in line with the actuality of the prior art.

 The objectives and results of the present invention are shown in Table 3. The resistance type along the way is a recognized target mode of diversion loss. The disclosed deflector in the following description can guarantee the realization of this mode, and it does not change the mode under variable operating conditions.

 The high potential ratio impeller can also have high impeller stroke hydraulic efficiency, so its overall hydraulic efficiency will reach a very high index. The substantial increase in hydraulic efficiency means that the first major bottleneck that restricts the efficiency of centrifugal pumps has been broken.

 In addition to improving the diversion efficiency, the solution of the invention can also greatly reduce the manufacturing cost of the centrifugal pump. The reasons are as follows: After the high potential ratio impeller is used to reduce the liquid flow speed, as the kinetic energy conversion load is reduced and the working speed is reduced, the flow of the deflector can be greatly shortened, the structure can be correspondingly cylindrical, and its volume can be greatly reduced. It is possible to reduce the manufacturing cost of the pump. In the prior art, the deflector works under the super-hydraulic specification condition of high-speed liquid flow, and the inlet linear velocity ranges from dozens of meters per second to tens of meters per second, which exceeds the hydraulic specification by an order of magnitude. Although it is impossible to achieve high kinetic energy conversion efficiency, there is still reason to dig out the efficiency potential of every percentage point as much as possible. However, under the existing technology framework, the space cost of improving the diversion efficiency is too great, and the potential is limited. We can see that the guide rings, guide wheels, worms and other deflectors arranged in the outer ring space of the impeller have a larger volume than the impeller due to their volume having a second-order function with the diameter dimension. The deflector has a larger volume ratio. They occupy most of the volume of the centrifugal pump, consume a lot of metal materials and processing hours, resulting in a lot of pump manufacturing costs. The solution of the invention reduces the kinetic energy conversion load of the deflector to a fraction, and reduces the working speed by about half. Therefore, the size and volume of the deflector can be greatly reduced, and the accuracy and surface quality requirements can be appropriately relaxed. The situation that the device occupies most of the volume will be greatly improved, and the manufacturing cost of the centrifugal pump is expected to be reduced accordingly.

The solution of the present invention also has a disadvantage, that is, the theoretical specific work and The theoretical head will decrease with the increase of the feedback reduction ratio K. This is because the output specific potential energy and output specific kinetic energy of the impeller decrease at the same time, which means that the output of the impeller with the same diameter and the same speed will be reduced. After the parameter K is selected, to achieve the same theoretical head, the diameter of the impeller or the speed must be increased to compensate for the loss of theoretical specific work and theoretical head. Obviously, this will increase the impeller linear velocity and the absolute velocity of the liquid flow, and The friction loss of the disk will be increased at a high power.

 This disadvantage may seem disturbing at first, but in-depth analysis shows that the loss caused by the disadvantage can only increase the friction of the disk, and the measures described below can solve this problem. Other seemingly unfavorable characteristics can actually be ruled out, and even such shortcomings can lead to encouraging new conclusions. The reasons are as follows: (1) The theoretical specific work and theoretical head have a square relationship with the involved speed, while the absolute speed and K have a linear relationship. Therefore, the theoretical increase in diameter or the product of diameter and rotational speed (inconvenient continuous increase in rotational speed) is required. The increment is small, and the required speed increase is much smaller than the absolute speed reduction caused by K. Second, due to the large increase in efficiency, the demand for theoretical specific work and theoretical lift has been greatly reduced, so that no compensation is actually required. It can also be seen from the analysis of the pressure coefficients to be discussed below that the theoretical pressure coefficient and the actual pressure coefficient with little change within the recommended K value range of the scheme of the present invention are basically equivalent to the design pressure coefficient commonly used in the prior art, which is approximately equal 1, so there is basically no need for compensation or the amount of compensation is small, which is the actual effect after considering efficiency factors. This conclusion can be understood as follows: The reduction of the theoretical specific work and theoretical head of the present invention actually reduces the part of the energy reserved in the prior art design that must be considered due to low efficiency, mainly the diversion hydraulic loss, which also includes Hydraulic loss of part impeller flow. Third, the reduction in theoretical specific work and theoretical lift and the reduction in parameter control is the basis point of the invention for changing the rigidity of the centrifugal pump parameters. It is from this base point that a wide range of adjustable and self-controllable designs has been created. space.

The conclusion is that the disadvantages and disadvantages of the solution of the present invention can be overcome and can be turned into advantages. The high potential ratio impeller scheme of the present invention includes the following specific designs, which can make the target performance more prominent, work more stable, and its design steps are more specific and specific:

 a. The outlet of the runner is rectangular, or inner chamfered rectangle, or circular, the perimeter is as small as possible, and its adjacent acceleration section has a corresponding section with continuous transition. When using internal chamfered rectangle or circular cross section, the cover may have corresponding supplementary structure. The angular distance between adjacent outlets is equal to 360 degrees divided by the number of channels, and the normal surface of the outlet is perpendicular to the channels. The connection from the inside edge of the previous exit to the outside edge of the next exit is a smooth involute arc cylinder, and the round or inner chamfered exits are connected by deep to shallow channels. The cross-sectional area of the channel between the cylindrical surface or the groove surface and the circumferential cylindrical surface has a periodic linear relationship with the center angle. The separately distributed exit stream bundles are sorted by the Coanda effect of the solitary cylinder or slot surface, forming an inwardly curved uniform flow line at the exit interval of the wheel. The radial and tangential components of the velocity do not change with the center angle. On a circular cylinder, the pressure, velocity, and their radial and tangential components are equal everywhere.

 b. The sum of the outlet area of each flow channel is equal to the ratio of the design volume flow rate and the design outlet relative speed, which is equal to the product of the impeller peripheral speed and the feedback reduction ratio K.

Technical element a in the specific scheme discloses a flow channel exit shape and wheel edge structure of a high-potential-ratio impeller and a mechanism for organizing the exit velocity field. Compared with the traditional impeller, this design and its mechanism increase the degree of restraint and stability of the outlet liquid flow, and its friction area is also small. The exit angle design reverses the exit relative speed to the implication speed and is designed to counteract that speed. The normal surface of the outlet perpendicular to the flow channel means that the flow line is smooth and the liquid flow is stable, which means that the outlet angle is small. The outlet angle sine determines the axial component of the relative velocity of the outlet, which is equal to this component divided by the relative velocity of the outlet. Due to the specific shape of the connection surface of the exit interval, the Coanda effect makes the axial surface velocity component of the flow field in the exit area uniformly distributed throughout the wheel circumference, so the axial surface velocity component is exactly equal to the flow divided by the cylindrical surface of the impeller exit Area. Its axial plane speed is small, and the exit angle sine is also small. The angle between the exit normal plane and the axial plane is also very small, almost on the axial plane. The cosine of the exit angle is almost equal to 1, so it has the greatest offsetting effect on the speed of implication. In addition to outputting high potential ratio liquid flow, the main consideration of this design is to stabilize the flow field in the exit zone. Stabilization goal This includes making the liquid flow neither locally excite the front area nor rim outflow at the exit; it also uses the Coanda effect to make it a unilateral rim constraint: it also includes the radial component of the exit velocity And tangential components have the best directional continuity and distribution uniformity.

 The connection between adjacent outlets by smooth involute arc cylinders or grooved cylinders is an important feature. The purpose is to make the exit stream beam uniformly absorbed and continuous by the Coanda effect of the contour cylinders or grooves. Bend inward to avoid outflow or local excitation in the longer exit interval area due to momentum inertia. The formation mechanism of the Coanda effect is: Under the condition that the neighborhood is fully constrained and within a certain range of momentum inertia, the difference between the absolute pressure of the outer area of the flow beam and the inner wall surface can provide a normal force to cause the liquid flow to continuously turn with the wall surface. The absolute pressure of the wall is adaptively changed with the external pressure. When the absolute pressure in the small neighborhood of the wall is adaptively reduced to equal to or close to the saturated air pressure, the Coanda effect is unstable and outflow occurs. The conditions for utilizing the Coanda effect in the present invention are: the outside of the exit stream beam has a sufficiently large static pressure guaranteed by the specific potential energy amplitude; the relative velocity of the liquid flow exit is restricted by the minimum potential-to-kinetic ratio, and its momentum inertia is limited; The surface is mathematically smooth, has continuity of curvature, and has a large radius of curvature. The required centripetal acceleration of the Coanda is always less than the amplitude limited by the specific potential energy. Therefore, the Coanda effect is stable and no outflow occurs.

The periodic linear relationship between the cross-sectional area of the flow channel between the outer cylindrical surface of the outlet or the groove surface and the circumferential cylindrical surface and the center angle is another important feature. Its role is to strengthen the Coanda effect, and the main purpose is to A uniform radial velocity component is regenerated in the exit interval. Since the outer side of the flow channel between the outer cylindrical surface of the outlet or the groove surface and the circumferential cylindrical surface is already open, the effect that the impeller can produce can only be a unilateral constraint of the Coanda effect. Relying on this constraint, a flow field in which pressure, flow velocity, and radial and tangential components are evenly distributed can be formed at the outlet and its interval position, which is an innovation of the present invention. The unilateral constraint virtually reconstructs the effect of continuous openings, and overcomes a series of problems such as rim backflow and outflow caused by rimless constraints in the prior art. This mechanism allows the internal and external The mouth area also satisfies the complete constraint conditions, so the problem of local excitation when the impeller flow channel and the deflector flow channel are connected is solved from the source. In addition, the uniform distribution of pressure and speed on the impeller's circumferential surface is a state independent of operating conditions, which results in complete adaptability to variable operating conditions.

 The large opening in the prior art is almost continuous. Due to the small relative speed and no radial restraint of the wheel in the radial direction, under the action of the abduction tendency of the relative vortex in the flow path of the finite blade impeller, a backflow, Many unstable phenomena such as turbulence and low pressure region at the tail of the suction surface. These phenomena are more harmful, and the hydraulic efficiency of the impeller is greatly reduced. This is a hot topic, and many researches have been conducted on it in the academic world, including the analysis results using three-dimensional turbulence theory and data measured using ultrasonic or laser technology. The literature is vast and the laws have been basically figured out. Under the guidance of the potential flow theory, many measures have been taken for the large recirculation of the pressure surface caused by the relative vortex. For example, a small back curve can slow down and control the recirculation, but it has increased the blade wrap angle and increased The cost of loss along the way. The vortex at the trailing edge of the suction surface of the blade of a large-open impeller is another important cause of the hydraulic loss of the impeller. This phenomenon can cause strong turbulence and deflow in the wake area when running under flow conditions, and may cause large turbulence in severe cases. Angle of exit congestion. The trailing edge vortex on the suction surface forms a suction obstacle to the rotating shaft. The force arm is long and the resistance torque is large, so the power loss is large. The trailing edge vortex is the result of the combined action of all complex force fields and cannot be explained by potential flow theory. Investigating its technical design reasons, the lack of constraints is definitely correct. The stability and uniformity of the large-open impeller flow channel and its exit flow field have seriously restricted the improvement of the hydraulic efficiency of the impeller process, and have not been solved for a long time. In fact, under the structural framework of the prior art, Among the constraints, it is possible to find the best compromise, but it is difficult or impossible to find a fundamental solution.

The solution of the present invention does not have a large opening, and a unilateral rim constraint is set in the exit area, which is a super-stability mechanism of the flow field in the exit area. Under this design, the relative vortex does not abduct, and there is no possibility of backflow at the exit wheel edge area, and turbulence and outflow in the wheel edge area cause damage. The instability phenomena of energy consumption and cavitation include the vortex loss at the trailing edge of the suction surface will no longer exist, the velocity in the exit zone will be stable, and the streamlines will be uniform, so higher hydraulic efficiency of the impeller can be expected.

 Technical element b provides a specific method for calculating the exit area and setting the feedback reduction ratio. The importance of the feedback reduction ratio parameter K has been mentioned before, and will be discussed repeatedly in subsequent descriptions. After the geometry and speed of the impeller are determined, parameters such as K and outlet area are determined according to the method described. In fact, the design conditions of the centrifugal pump have been fully determined, and its various characteristic curves will be determined accordingly.

 The specific design also determines the design condition parameters of the impeller at the same time. When the operating conditions change, the feedback reduction ratio K changes proportionally to the flow rate, and the specific work or head has a linear decreasing function with the flow rate. These relationships also show that the product of the present invention is capable of operating under varying conditions over a wide range. The operation under variable working conditions is a broad requirement. The present invention is different from the operating characteristics of the prior art and obviously meets the actual needs of operation under variable working conditions. For example, the unique flow-power characteristics of the present invention reflect better adaptability to changing working conditions: the shaft power increase is significantly smaller when the flow is increased, and there is no danger of overload; when the flow is reduced, its head is sharp As the power is increased, the magnitude of the power reduction is also small. When the head is increased to the maximum, the power decreases with the decrease of the load and approaches the fixed mechanical power loss. This is an ideal characteristic that the low end varies with load and the high end approximates constant power. As another example, the flow-efficiency characteristic peculiar to the present invention can better reflect its adaptability to changing working conditions. In an externally regulated liquid flow system that is usually operated under-flow, the characteristic of the low-end efficiency of the prior art centrifugal pumps, which causes a steep drop in efficiency, will result in serious energy waste. However, the efficiency characteristic of the present invention is increased at the low end. After the internal friction reduction technology described below is adopted, the area where such low end efficiency does not decrease but rises can even be shifted to the lower end by an index on the logarithmic coordinates. Interval.

The blade shape design scheme of the high-potential-ratio impeller of the present invention is: the impeller blade is L-shaped, and the front and middle portions thereof are straight line segments, which are in a radial direction, and the elbow and the tail portion are bent to the anti-tangential direction with an appropriate curvature radius transition, and the tail portion has Isolate the mechanical strength and sharp end of the internal and external pressure difference, the transition of the proper curvature radius includes the shape change of the internal and external sides, the tail The distance between the inner side of the part as the outer restraint edge of the acceleration section and the blade elbow meets the change requirements of the cross section of the acceleration section of the flow channel, the outer side of the tail meets the change of the strike angle of the Coanda effect finishing, and the outer radius of curvature of the elbow also meets the no-flow condition.

 The above-mentioned blade shape scheme gives the specific structure to realize the wheel edge restraint and the blade groove tail acceleration. The rim restraint is achieved by the near-circumferential trend of the tail of the L-shaped blade. The inner side constitutes the outer restraint wall surface of the tail end acceleration section of the blade groove, and the outer side constitutes the involute contour cylinder surface or groove surface outside the outlet, which is used to restrain the outside of the mouth. Flow field. It is very convenient to use the numerical algorithm to construct the inner and outer marginal curves of the blade. Using this method to calculate from the circumference to the inside, the functional data such as the positioning of the contour line outside the mouth, the thickness of the blade, and the cross-sectional area of the blade groove can be realized point by point. Perform mechanical calculations such as force, strength, and fluid motion. Among them, only the marginal line on both sides of the tail and the pressure side curvature of the elbow are hydrodynamically sensitive. There are two main requirements. One is to gradually reduce the cross-sectional area of the blade groove to reach the design area by the exit; the other is to have good mechanical properties. The intermediate process has infinitely many possible choices. Among them, the optimal design exists and can be found, but the basic requirement is marginal smoothness. Guaranteeing the mathematical smoothness of the margin line is a requirement for the first-order spatial change rate, and the corresponding mechanical value is the directional continuity of the velocity. The test of non-diverging mechanics equations mainly uses the Newton's law of particle curve motion and the pressure difference check as the source of power, which can generally be satisfied.

 The above detailed schemes are quite different from the prior art, but they are reasonable. First, the groove width of the impeller with a finite number of blades is geometrically limited, and is proportional to the sine of the blade's heading angle. The front middle part of the blade is designed as a straight line segment and a large radius of curvature and has a radial trend. Its groove and its entrance The cross-sectional area is maximized and its relative flow rate is thus minimized. Second, the L-shaped blade is an inevitable or optimized choice after reducing the outlet area of the flow channel. Because the distribution of the small outlets on the perimeter of the impeller is necessarily separated from each other, this requires an arc structure on the rim to construct this separation, and is used as the isolation and pressure of the blade groove flow channel and the outer exit area of the wheel. Structure, this structure can only be the curved blade tail.

This improved solution of the present invention reduces the cross-sectional area of the blade groove while reducing The blade wrap angle reduces the groove process. Compared with the prior art backward curved blade design, the cross-sectional area of the blade groove loading area is approximately doubled, and the flow process of the blade groove loading work area is reduced by approximately 50%. According to this rough estimate, compared with the traditional centrifugal pump impeller with the same diameter, speed and flow rate, the flow velocity in the loading area of the blade groove of the solution of the present invention will be reduced by 50%, and the friction loss per unit length will be reduced by 75%. Friction losses along the slot loading zone will be reduced by 87.5%.

 The tail of this improved solution is the acceleration zone, and its cross-sectional area is continuously and rapidly reduced at an appropriate rate of change, and the flow velocity is therefore rapidly and continuously increased. This is an energy conversion process in a rotating coordinate system, and certain losses will occur. . When designing the accelerating flow path specifically, the exit cross-sectional area determined by the flow rate and the relative velocity of the exit should be used to determine an appropriate cross-sectional area change rate to infer the size. The spatial change rate of the flow velocity and the shape of the blade groove will be determined from it. Because this process segment is short, when K = 0.5, the hydraulic loss is not significant. The basic equation of the acceleration section is: specific kinetic energy increase = specific potential energy reduction X efficiency, where the efficiency factor is related to the cross-sectional area change rate and surface quality and other factors. With reference to empirical data analysis of similar jet technologies, its acceleration efficiency can usually be It reaches 98%. Therefore, according to the analysis of potential flow theory, the total hydraulic loss of the L-shaped blade flow channel will be significantly reduced compared with the prior art. In fact, the further comparative advantage will mainly come from the differences in the stability of the flow field, the unstable phenomena such as turbulence, backflow, and wake vortex caused by the relative vortex abduction and secondary flow caused by the large opening. In the prior art, the hydraulic loss of the impeller runner is relatively large. In comparison, the flow field of the L-shaped blade solution of the present invention has less instability factors. Except for the relative vortex effect in the blade groove, there is no outlet return flow and trailing edge vortex, and the hydraulic efficiency of the impeller will be greatly improved. Moreover, the technical features described later in the present invention can basically restrain the relative vortex in the blade groove, and its efficiency will be more advantageous.

The radial trend of the front and middle part of the blade in the improved scheme of the present invention is also significant for improving the cavitation resistance. In the prior art, the areas with the most severe cavitation damage are the sides of the blades in the inlet area and the suction side in the outlet area. For the entrance zone, the leaves of the invention The chip design pitch is doubled, and its flow rate will be halved at the same design flow. In combination with the adaptive pre-rotator described in the following description, this impeller will have particularly good anti-cavitation characteristics in the inlet area. For the exit area, due to complete constraints, the trailing edge vortex or the low-pressure area of the suction surface no longer exists, and there is no possibility of cavitation under higher positive pressure.

 The design of the L-shaped blade in the radial direction of the front part of the present invention and a 90-degree inlet angle is a special design suitable for high-potential centrifugal pumps, which has a small blade wrap angle, a large cross-sectional area in the front part of the blade groove, short flow, and flow velocity. The advantage of low friction loss is small. For low specific speed impellers, the advantages will be more obvious. In addition to the main target characteristics of outputting high potential-to-dynamic-ratio liquid, this impeller also has the advantages of high hydraulic efficiency and good anti-cavitation characteristics of the impeller flow.

 The invention also includes a matching accessory design, that is, an axial or radial inflow adaptive pre-spinner that rotates coaxially with the impeller is installed at the impeller suction chamber or the front stage deflector outlet. It consists of blades, wheels and rigid ribs. The number of blades is less than the number of blades of the impeller. The blades are made of composite materials. It has a tensile coefficient of elasticity that gradually increases from the front end to the root. In the assembled position, the wheel hub is freely sleeved on the rotating shaft or the impeller shaft sleeve, and the front end of the blade is suspended on a rigid rib at the entrance, and a pre-spinning flow path is formed between the blades. Among them, the rigid ribs of the axial flow pre-rotator are arranged at a radial position on the circumferential surface of the inlet, and the rigid ribs of the radial flow pre-rotator are arranged at a position parallel to the rotation axis on the cylindrical surface of the inlet. During operation, the elastic sail blade will deform into an adaptive flow channel with the change of the fluid flow parameters, and its angle of attack and inclination along the way are adaptively changed.

The prior art centrifugal pump design concept believes that positive pre-spin should be avoided, and its guiding idea is to increase the specific work and head by increasing the relative speed of the inlet. For example, the anti-lead vane exit angle in the prior art multi-stage pump design specifications is designed not only to eliminate the original loop volume, but also to apply a reverse loop volume, obviously for this purpose. As a result, the lift is increased, but the price paid is increased inlet hydraulic loss, poor cavitation resistance characteristics, and poor operating adaptability under variable operating conditions. The invention is based on a conservative loop design The idea is to maintain a conservative or positive spin in the forward bearing neighborhood of the impeller flow path entrance. The design of the adaptive pre-rotator is the product of this guiding idea, which is the opposite of traditional design and the effect is exactly the opposite. The role of the pre-rotator is to reduce the specific work of the impeller that should not have been increased, in exchange for small inlet hydraulic losses, good anti-cavitation characteristics, and the benefits of adaptive adaptation to variable operating conditions. It is a kind of technical trade-off between small and big. In addition, the aforementioned characteristic design of the 90-degree inlet flow channel also needs a pre-spinner to be matched.

 The adaptive pre-spinner is a separate regulating work component, which is equivalent to a special small impeller, which is used to adjust the size and distribution of the incoming flow velocity to meet the needs of the impeller suction. When the rotational energy of the incoming stream is insufficient, the pre-rotator output specific work is increased. When the incoming flow rotation speed is too large, the pre-spinner can also absorb many parts in the speed field arrangement, and it will change into the working state of the water turbine itself. Energy adjustment is a macro statistical effect of the velocity field arrangement of the incoming flow. The meaning of the velocity field arrangement is: distributed energy exchange through the elastic propeller-shaped flow channel, and the spatial distribution of the magnitude and direction of the incoming flow is continuously changed. It is adapted to the velocity distribution determined by the overall operating conditions at the entrance of the blade groove.

These features of the prespinner are mainly to match the radial direction of the front of the L-shaped blade, and at the same time to adapt the impeller to the needs of variable operating conditions. The mechanism to achieve these functions is: When the working condition parameters change, the change in the suction chamber area is reflected in the change in flow rate and the change in the magnitude and direction of the speed caused by the change in flow rate. Because the pre-spinner and the impeller rotate coaxially, the rigid ribs at the entrance and the hanging sail blades rotate at a hook speed, and the speed at which they cut into the flow will change with the change in flow. In a spatial neighborhood of the cut-in line, the inertial force of the liquid flow and the distribution of the tension at the entrance of the soft blade will reach a balance at each point on the blade. The geometric effect of this force balance point is the sail blade The beginning will adaptively change the angle of attack as the incoming flow speed changes, thereby automatically maintaining a state tangent to the incoming flow line. This mechanism solves the problem that the change in the magnitude and direction of the speed of the traditional impeller when the flow changes will result in a fixed angle. Fixed blades collide, which causes problems such as inlet turbulence and unsuitable changes in working conditions. After the incoming flow enters without impact tangent in the inlet area, the state of tangential flow can be maintained in the subsequent pre-swirling flow channel area. The mechanical principle is still the balance of the inertial force of the flow, the blade tension, and the bending resistance. Among them, the larger the radial bending strength of the blade will gradually increase the normal restraint strength until the exit connected to the impeller flow channel, the exit direction will basically be the impeller inlet direction. The incoming flow is continuously changed in speed and direction by the blade restraining force in this flexible and gradually decreasing propeller-shaped flow channel. When it arrives at the exit, it has a preset direction, and the speed of the flow channel is It can change with the change of the flow rate, and the change of the speed of each place under the same flow rate is the result of the blade constraint and the energy exchange due to the constraint.

 The pre-spinner makes the impeller flow line in the inlet area uniform and stable during variable flow operation, without impact, turbulence, and flow loss. This keeps the hydraulic loss in the inlet area low and greatly improves cavitation characteristics.

 The series design of the present invention also includes another more important feature, that is: a uniform speed bifurcation to suppress relative vortex is arranged in the impeller groove. Each blade channel is divided longitudinally by 1 to 3 uniform-speed combing leaves, forming 2 to 4 bifurcations. The entrance of the fork is close to the entrance of the slot, and its cross-sectional area is evenly distributed. The exit of the fork is close to the exit of the vane slot, and the cross-sectional area is determined based on the superimposed relative vortex dynamic distribution and given speed distribution, or by empirical data optimization optimized by preferred experiments. The uniform-speed combing blade exerts a resistance and restraint force on the relative vortex, forms a uniform-velocity of the groove speed, and generates the required inlet pressure gradient and outlet speed gradient.

The blade groove average speed bifurcation of the present invention is a special structure specially designed to solve the key problem of relative vortex which affects the hydraulic efficiency of the impeller. Relative vortex is the inertial motion of free fluid inside a finite-blade number of impellers. This motion will produce a different velocity field structure than the infinite-blade number of impellers and change its specific work value, reducing the theoretical head of the pump by 15% to 25%. In addition, the relative vortex also directly generates serious losses, and its velocity and pressure are unevenly distributed. It superimposes on the flow field of the lobes, forming backflow, turbulence and In the low-pressure deflow zone, large losses are generated, which further reduces the actual head. The superimposed relative vortex of the liquid flow will generate an additional pressure difference between the pressure surface and the suction surface, which will directly form an additional resistance torque and an additional power loss, which will increase the proportion of the invalid specific work, which is one of the reasons for the further reduction of the actual head.

 The seriousness of the problem can also be seen in the design practice of the pressure coefficient method of the centrifugal pump. The theoretical pressure coefficient of an impeller with infinite blades without relative vortex is equal to 2, where the specific potential energy and specific kinetic energy partial pressure coefficients are 1 respectively. However, the pressure coefficient of the traditional centrifugal pump design usually only reaches about 1, and the highest is only 1.1. The reduction of the pressure coefficient is above 0.9. There are three main reasons for this: the finite blade reduces the specific power of the transmission and the theoretical lift; the impeller flow loss reduces the actual head; the diversion flow loss further reduces the actual head. The first two of the three main causes are relative vortices, which shows the technical importance of curbing relative vortexes.

The intensity of the relative vortex increases monotonically as the number of blades decreases. For a conventionally selected ratio scale with the number of blades equal to 6 to 8, the absolute value of the relative vortex theoretical angular velocity is almost equal to the impeller angular velocity, and the maximum value of the theoretical linear velocity is about 1/3 of the peripheral velocity of the impeller. May reach a dozen meters per second. Such a large vortex speed and a small relative speed for vector synthesis, for a traditional impeller with a large liquid flow exit angle, will generate a high-speed forward flow on the suction surface with an absolute value 1 times the relative velocity of the edge vortex speed. Produces a pressure surface backflow with an absolute value that is 1 times lower than the edge vortex speed but still a high velocity. Although marginal friction and internal viscous forces will curb this high-speed flow and reduce it to a lower equilibrium value, losses will occur. When the problems of backflow, turbulence and deflow caused by relative vortices are serious, the impeller's design outlet will be squeezed up, which will become a harmful space with excessive loss, and it may also generate pressure wave of blade frequency and generate radial vibration of hundred hertz level. Of course, for the present invention, the outlet backflow has been completely eliminated, and most of the above-mentioned problems have been solved, but the relative vortex inside the blade groove flow channel still exists. Therefore, a method of restraining relative vortex is adopted instead of resistance, which is used to apply targeted shear moment to make the blade attract The deceleration of the force surface and the acceleration of the pressure surface, while absorbing and transforming the remaining vortex energy, have become a necessary improved design of the present invention.

 The uniform speed bifurcation of the present invention is the product of the above design idea. According to the design requirements, the average speed bifurcation is arranged at a wide cross section in the middle of the front of the chute flow channel, and the crowding coefficient of the cross-sectional area of the convection channel is not large. From the foregoing analysis, it can be known that the velocity of the wide area of the blade groove is originally low. If the relative vortex is restrained, the positive superposition of the vortex speed on the suction surface and the negative superposition of the pressure surface will be eliminated, and the blade groove will return to normal. Low-speed flow. The principle of curbing relative vortexes of the present invention is as follows: 1) The binding forces of the pressure surface and the suction surface of a branch road in a generally radial direction are unevenly distributed. The shearing moment is equivalent to increasing the number of blades. 2) The bending cooperation between the L-shaped blade and the fork structure makes the inside and outside walls of the fork unequal in length, and the internal and external forks are unequal in length and have a large difference. The requirements of the vortex resistance are consistent; 3), more importantly, the uneven distribution of the cross-sectional area of the branch exit controls the internal and external flow velocity of each basin, so that the branch near the suction surface of the main blade has a higher exit speed and is closer. The branch road on the pressure surface has a lower exit velocity. The difference in the reaction resistance will produce a restraining torque on the circumferential surface relative to the center of the entire vortex vortex. Or, the difference in the reaction resistance is exactly equal to the vortex involved in the branch. The power difference makes it possible to achieve the desired uniform speed distribution.

Because the bifurcation is distributed in the non-import and export areas of the leaf tips, the remaining import and export areas will still have a relatively small vortex tendency. The solution requires that the ratio of the cross-sectional area of the exit of the average speed fork should generate an appropriate inlet pressure gradient to generate the shear moment that curbs the vortex in the entrance zone. It should also generate an appropriate exit velocity gradient to change the pressure at the outlet of the fork Distribution, injecting a bifurcation near the pressure surface, and reducing wheel friction around the exit area of the blade groove. The detailed and accurate dynamic calculation or asymptotic test can achieve the appropriateness as described, and the results will be implemented into the specific specific size of the branch structure and optimized. It should be noted that most of the vortex power is offset by the reaction resistance when the acceleration pressure is formed in the near-suction bifurcation. There is no positive vortex energy accumulation in the rotating coordinate system. The problem is that only the liquid flow flowing out in the positive direction will be affected by the vortex dynamics, and the specific energy increase will form the velocity gradient.

 The average speed fork is a very important innovation. Proper branch exit area ratio can achieve resistance balance in two dimensions in the blade groove, which can almost completely restrain the relative vortex in the impeller flow channel, and the hydraulic efficiency of the impeller flow channel will be greatly improved. Erosion characteristics will also be greatly improved. The effective containment of the average vortex by the relative vortex will make the performance of the impeller close to the performance of the infinite number of impellers. The most important contribution includes the recovery of the theoretical Yang Cheng and the reduction of hydraulic loss due to the impeller flow. The difference between theoretical head and actual head is reduced.

 Because the theoretical specific work and theoretical head of L-shaped blade impeller with average speed bifurcation are very close to the law of Euler equation of infinite blade number, this equation can be used to discuss the performance of high potential ratio impeller with limited blade number.

Under the condition that the inlet angle (relative flow angle) of the L-shaped blade impeller flow channel is equal to 90 degrees and the outlet angle is approximately equal to 0 degrees, according to the inlet and outlet velocity triangles, there are ^ ²

(u ₂ -w ₂ ) ² . Reintroducing the impeller inlet implication speed coefficient χ = Ui / u ₂ and inlet relative speed coefficient μ = Wi / w ₂ , then there are U! ² = u ₂ ¾ ² , W! ² = u ₂ V ^{2 2} , w ₂ ² = u ₂ ² K ² , v ₂ ² = u ₂ ² (1-K) ² and _Vl ² = u ₂ ² (χ ² + μ ² Κ ² ). All terms are normalized to u ₂ , and the set parameter form of the equation is as shown in formula (3), which contains the theoretical pressure coefficient ψτ which is of great significance, and its expression is shown in formula (4).

^{Υτ = 0 · 5ω 2 ((} 1 - χ 2Κ 2 (1 - μ 2)) + (1 - χ 2 - 2Κ + Κ 2 (1-μ 2))

= ω¾ ² (1-χ ² -Κ) (3)

^{ψτ = (1-χ 2Κ 2} (1- μ 2)) + (1 - χ 2 - 2Κ + Κ 2 (1 - μ 2))

= 2 (1-χ ² -Κ) (4)

In both formulas, (1- χ ² -Κ ² (1- μ ² )) is the specific potential energy incremental pressure coefficient, and (1-χ ² -2κ + κ ² (1- μ ² )) is the specific kinetic energy incremental pressure. Coefficient, κ is feedback deceleration The ratio, ω is the impeller angular velocity, and R ₂ is the impeller radius. Because the value of the inlet implication speed coefficient χ equals the ratio of the impeller inlet and outlet radius is a fixed geometric parameter. According to the characteristics of the incompressible fluid, the value of the relative velocity coefficient μ- at the inlet is equal to the inverse ratio S ₂ / S of the cross-sectional area of the inlet and outlet of the impeller flow path, which is also a fixed geometric parameter. After the parameter transformation, the equation only includes the peripheral speed of the impeller and three normalized coefficients, which greatly facilitates the analysis and discussion of the characteristics of the new impeller.

In equation (3) and theoretical pressure coefficient expression (4), there are various needs for retaining the expressions of specific potential energy increment and specific kinetic energy increment partial pressure coefficient. There are two kinds of theoretical analysis, theoretical and actual potential kinetic energy production ratios. They are all used in the dependent analysis. Obviously, the composition of the two terms of the two expressions highlights the mechanical sources of the specific potential energy incremental partial pressure and the specific kinetic energy incremental partial pressure and their proportional relationship. Among them, the specific potential energy incremental pressure coefficient is composed of the algebraic sum of the centrifugal force work pressure coefficient 1-χ ² and the potential energy consumption pressure coefficient decrement term-κ ² (1- μ ² ) for accelerating relative motion, and the specific kinetic energy The incremental pressure coefficient is the pressure coefficient decrement term of the residual acceleration force work pressure coefficient 1-χ ² of the blade equal to the centrifugal force work pressure coefficient, and the kinetic energy is fed back to the rotating shaft to reduce the work force-2K and increase due to relative motion acceleration Algebra and composition of the pressure coefficient incremental term K ² (1-μ ² ). -The existence of the terms K ² (l-μ ² ) and Κ ² (1-μ ² ) indicates that the process of converting potential energy to kinetic energy exists when the flow at the tail end of the channel is accelerated. Positive correlation and μ negative correlation have sharp changes, but they cancel each other out in the overall theoretical pressure coefficient without affecting. From the above analysis, we can see that the centrifugal force work pressure coefficient 1-χ ² and the blade residual acceleration force work pressure coefficient 1-χ ² are both equal, which is the general law of centrifugal pump impeller specific work distribution and head generation. When no technical treatment is performed, the theoretical pressure coefficient of the centrifugal pump impeller will be 2 (1- χ ² ), and the theoretical potential-to-moment ratio is equal to 1. Obviously, two 1-χ ² terms are also generated in the present invention. Under the action mechanism of the kinetic energy feedback technology, the impeller feeds back the specific kinetic energy with a normalized coefficient of -2K to the rotating shaft, and the theoretical pressure coefficient is reduced to 2 ( 1- χ ² -Κ), and the theoretical momentum ratio is thus large. Increase in amplitude. Because the potential and kinetic energy production have different actual conversion efficiencies, the actual pressure coefficients need to be calculated according to the hydraulic efficiency of the respective processes when accounting for losses. At this time, the theoretical pressure coefficients of specific potential energy increase and specific kinetic energy increase are An intermediate parameter that must be used.

 By carefully analyzing the dependent law of the theoretical pressure coefficient, you can theoretically see the advantages of the blade structure scheme of the present invention in improving efficiency and improving operating characteristics, and then you can find its great potential in controlling specific power output. This discovery will become the theoretical basis for designing adjustable centrifugal pumps and self-controlling intelligent centrifugal pumps.

 Equation (3) shows that in the composition of the theoretical pressure coefficient of the present invention, the specific potential energy incremental pressure coefficient occupies the majority, and the value of the specific kinetic energy pressure coefficient is relatively small, and due to the large decrease in absolute speed, The hydraulic efficiency of the diversion process is greatly improved, so the actual pressure coefficient of the pump is very close to the theoretical pressure coefficient. Based on these two reasons, the latter can basically be used instead of the former when designing the impeller. If accurate calculations are required, the diversion efficiency formula can be applied or further modified using the full-range hydraulic efficiency formula. The variation of the theoretical pressure coefficient of the high potential ratio centrifugal pump with χ and κ is shown in Table 4.

High potential ratio centrifugal pump theoretical pressure coefficient ψ = 2 (1-χ ² -Κ)

Referring to Table 4, the theoretical pressure coefficient decreases in a square relationship as the implication speed coefficient χ increases, and decreases linearly as the feedback reduction coefficient K increases. Both It is a true decimal, so the latter has higher sensitivity, which makes the high potential ratio centrifugal pump have approximately constant power characteristics and good linear throttling characteristics of adaptive power adjustment. χ is a design geometric parameter, which is limited by the flow rate, inlet flow velocity, and shaft diameter, and its value is usually about 0.2 to 0.3, which has little effect on the theoretical pressure coefficient.

 The theoretical momentum ratio of the impeller having the foregoing technical features of the present invention is:

^{_{λ τ = (1 - χ 2Κ}} 2 (1 - μ 2)) / (1 - χ 2 - 2Κ + Κ 2 (1 - μ ") (5)

Because the hydraulic efficiency of the impeller flow section of the present invention is very high, the actual output potential-to-dynamic ratio will be very close to the theoretical potential-to-dynamic ratio. The difference between the two mainly comes from the influence of the loss at the entrance and the acceleration stage. The entrance loss affects the specific potential energy of the output, and both of the acceleration stage losses are affected. Since such losses have been reduced very low, they can usually be ignored. in! ^: In the actual value range of μ, χ, the value range that λ _τ can reach is about 3-9. The physical significance is that the pressure potential energy increase produced by the high potential ratio impeller will be 3-9 times the kinetic energy increase reflected by the absolute velocity of the liquid flow. It can be seen that the high potential ratio characteristic of the present invention is very significant.

Specific kinetic energy increase and specific kinetic energy are not the same concept. When the object under investigation is an impeller, the intervention of external energy such as the pre-spinner and the external driving speed of the inlet flow requires the two concepts to be used differently. When the object to be investigated is the entire centrifugal pump, the two concepts can be regarded as the same except that there is an externally driven pipe inlet speed. In addition, when the diversion load is concerned, it is required to examine the specific kinetic energy increase, and when the diversion loss is concerned, the specific kinetic energy is required to be examined. In fact, the difference between the normalization coefficients of the two is only in the high-order small-quantity term μ ² κ ² , and its value is obviously very small, which can be ignored in general analysis.

² = 0.1 when taking the typical parameters μ, ψ kinetic pressure coefficient than _{2, = (1 - Κ)} 2, than the potential incremental pressure coefficient ^{11 ^ = (1 - χ 2} - 0.9Κ 2), specific kinetic energy increase in pressure Coefficient ψ ₂

Their changes are listed in Table 5. Table 5 Specific kinetic energy pressure coefficient, specific potential energy incremental pressure coefficient, specific kinetic energy incremental pressure coefficient, and potential-to-kinetic ratio change table

It can be seen from Table 5 that the two partial pressure coefficients decrease with the increase of χ and K, but there is a significant difference in the rate of change, and this difference causes the momentum ratio to increase with the increase of χ and κ. As the value increases, the change with χ increases is relatively slow, but the change with κ increases is very sensitive, which is exactly what the actual design needs. It is obvious that κ as an adjustment amount equivalent to χ ² should have high sensitivity, and κ is also different from χ, which is a geometric parameter that is restricted by the structure, and can enjoy a relatively large setting independence and flexibility. The value of K The range is usually about 0.5.

Theoretical pressure coefficient ψ = 2 (1- χ ² -κ), specific potential energy incremental pressure coefficient _Ψι = 1-χ ² -Κ ² (1-μ ² ), and specific kinetic energy incremental pressure coefficient ψ ₂ = 1- χ ^{In 2} -2Κ + Κ ² (l-μ ² ), (1-χ ² ) is used as the centrifugal force work or the remaining acceleration force work pressure coefficient term. They are the sole or first source of the two partial pressure coefficients, respectively. Changes are shown in Table 6.

 Table 6 Equivalent changes in the pressure coefficient of centrifugal force work and the pressure coefficient of remaining acceleration force

As the sole source of the specific potential energy incremental pressure coefficient, the pressure coefficient of centrifugal work given in Table 6 is obviously the upper limit of this coefficient. As the first source of the specific kinetic energy incremental pressure coefficient, the residual acceleration force work pressure coefficient given in Table 6 constitutes the main positive part of the coefficient, especially when χ and K are small. Taken together, as mentioned earlier, the two pressure coefficients given in Table 6 are two equal energy energetic springs of the theoretical pressure coefficient of the centrifugal pump. Their sum determines the upper limit of the theoretical pressure coefficient of the pump. When there is a typical parameter χ = 0 · 2 ~ 0.3, this upper limit is about 1.92 ~ 1.82. Considering that the design pressure coefficient of the prior art is usually around 1.0. Therefore, when the impeller of the present invention with the same wheel diameter and the same speed is required to achieve the same design pressure coefficient, it also has a pressure coefficient space of about 0.92 to 0.82 for setting the feedback reduction ratio κ and planning a greatly reduced pressure. Coefficient loss. At this time, the momentum ratio can reach about 3 to 5, which has basically met the needs. If it is necessary to further increase the potential-to-dynamic ratio, the design value of the pressure coefficient can be appropriately reduced, and the design lift can be achieved by increasing the rotational speed or appropriately increasing the wheel diameter. Because the related power of the friction loss of the former is relatively low, the former should usually be preferred. However, after using the internal friction reduction technical features described later in the present invention, the friction loss of the disk will become a problem that does not require special consideration. At this time, the pressure system is boldly reduced. By increasing the number and momentum ratio, higher hydraulic efficiency and overall efficiency can be obtained.

 In summary, the present invention improves the impeller of the prior art centrifugal pump relatively thoroughly. The main technical features include:

 1) The impeller flow channel adopts a small anti-tangential outlet that is separated from each other to generate a kinetic energy feedback deceleration mechanism and increase the resistance-to-potential ratio;

 2) L-shaped blades are used, whose front end is a radial groove, with a large cross-sectional area and a small wrap angle, a low flow velocity and a short flow, and a trailing edge restraint effect at the tail, which completely eliminates the phenomenon of backflow and outflow;

 3) The uniform speed bifurcation resistance is set in the blade groove to eliminate the relative vortex of the blade groove, so that the velocity field of the blade groove flow channel is evenly distributed, and the unstable phenomena such as turbulence, trailing edge vorticity, and turbulence between the pressure surface and the suction surface are eliminated. Resistive pressure difference.

 These improvements can achieve the following functional or performance characteristics:

 1). Decrease the speed of guide, reduce the loss of diversion, basically eliminate the loss of impeller, and greatly improve the hydraulic efficiency of the whole process;

 2) With the adaptability for variable operating conditions, the hydraulic efficiency of small flow operation is not reduced but increased;

 3) Increase the theoretical specific work and theoretical head of the impeller to make it closer to the law of the number Euler equation of infinite blades;

 4), reduce the pressure coefficient loss of the impeller process and the diversion process to a small, so that the actual head is close to the theoretical head;

 5). Insert the feedback reduction coefficient K which can be set flexibly into the basic equation, so that the theoretical head, actual head and hydraulic efficiency all monotonously change with the feedback reduction coefficient K, thereby establishing the adjustable and self-controlling centrifugation. The cornerstone of pump technology.

The present invention also pays great attention to the problem of disc friction loss and large and unstable axial thrust that cause the main loss of mechanical efficiency in the centrifugal pump. A technical solution to achieve the second purpose of the invention is designed: a closed impeller The end cavities on both sides of the disc or the rear cover side end cavities of the semi-open impeller are placed in the gas circulation or gas-liquid two-phase flow cycle The end cavity is filled with insoluble gas, and the impeller disk rotates in the gas phase medium, and its friction loss is small. The pressure of the end cavity gas dynamically maintains a balance with the surface pressure of the rotating liquid flow along the edge of the end cavity during the cycle, and is equal to or less than the impeller output. Static pressure, when the front and rear chambers are inflated, the specific pressure of the two is equal or similar. The axial force exerted by the gas on the impeller is equal to the product of the gas pressure and the area of the end face of the impeller, and has nothing to do with the size of the leakage gap and the leakage flow.

 Disk friction is a fixed loss that is not directly related to head and flow. Its magnitude is directly proportional to the 5th power of the impeller diameter and proportional to the 3rd power of the rotational speed. The loss of efficiency caused cannot be ignored. For low specific speeds, high heads, and small flow operations that deviate from design conditions, the relative impact of wheel friction is particularly serious. Numerical analysis shows that this type of loss is one of the main reasons for the low design efficiency of low-specific-speed centrifugal pumps, especially high-lift centrifugal pumps with relatively low speeds and large impellers. In addition, this kind of loss is also the decisive reason for the serious decrease in efficiency when the centrifugal pump deviates from the design condition at low flow rate operation, because the impact of the decrease in hydraulic efficiency on the overall efficiency has an asymptotic lower limit determined by the momentum ratio, and the wheel disk The ratio of friction loss to effective shaft work can be reversed to greater than 1 or even greater than 1 without restriction.

As far as the axial thrust is concerned, the mechanism of the traditional axial impeller forming the thrust on the pressure side is mainly due to the uneven effect of the output pressure on the end cavities on both sides, while the compensation of the reaction force for the change in the flow volume of the suction side is relatively small , And changes with flow. The reason for the difference in end-cavity pressure is that the average specific pressure of the end-cavity on the suction side is lower than the pressure side and its area is smaller. The difference in specific pressure on both sides is caused by factors such as the difference in gap width, the direction of the leakage flow and the difference in the momentum moment it carries, and the difference in the size of the centrifugal force field. Except for the area and the direction of leakage flow, these factors are unsteady and have a large proportion of change. Therefore, the traditional impeller will generate a large amount of non-smooth axial thrust. This thrust will lead to increased bearing losses and reduced efficiency, and may even cause mechanical failure. The problem of axial force balance increases the difficulty of designing and manufacturing a centrifugal pump. Specially designed balance plate device increases structural complexity and shaft precision The requirement of degree increases the cost of the pump, and the diversion effect of the balance disk will cause the volumetric efficiency and the total efficiency to decrease.

 The effect of the internal friction reduction scheme of the present invention is twofold-it can eliminate most of the disk friction losses, improve the internal mechanical efficiency of the pump, and at the same time reduce and stabilize the axial thrust. The principle of friction reduction is easy to understand. Since the viscosity coefficient of gas is two orders of magnitude smaller than that of liquid, the friction loss of the disc end face is reduced by two orders of magnitude. However, in practical applications, due to the random perturbation of the gas-liquid interface (liquid level) of the gas-filled end-cavity, bubble escape occurs at the trough of the perturbation wave, so the gas cannot completely fill the entire end-cavity. In the presence of an annular liquid region outside the end cavity, it is difficult to reduce the friction loss of the disc by two orders of magnitude. Based on this, the present invention describes the friction reduction effect by using the ratio of the friction loss of the disc with a residual annular liquid phase region when inflated to the friction loss of the disc when not inflated (tube friction loss Lg), ignoring the gas phase friction in the inflation region This ratio is given by equation (6).

Zj / Za = J _dl ^d W27rrdr / J ₀ ^d2 or ³ 27irdr = (d ₂ ⁵ -(^ ⁵ ) / (1 ₂ ⁵ = 1-(dx / dz) ⁵ (6) friction loss ratio / Z ₂ is the end As a function of cavity inflation diameter ratio (M ₂ ), the former is a measure of the friction reduction effect when there is a disturbance, and the latter is a direct measure of the inflation effect when there is a disturbance. The friction loss ratio calculated according to formula (6) is shown in Table 7 Data sheet friction loss ratio when the end cavity is not fully inflated

It can be seen from Table 7 that when the inflation diameter ratio reaches more than 96%, the wheel friction loss ratio will be reduced to less than 18%, and when the inflation diameter ratio reaches 99%, the wheel friction The friction loss ratio will be reduced to 5%. It can be seen that the actual friction reduction effect is to reduce the friction loss by an order of magnitude, rather than two orders of magnitude.

 The following specifically analyzes the quantitative relationship between the loss reduced by the internal friction reduction design of the present invention and the efficiency index of the centrifugal pump. The efficiency formula of the centrifugal pump containing the classic concept is shown in (7)

Where η,, and are the total efficiency, hydraulic efficiency and volumetric efficiency, respectively, (1-P _r Pi) is the internal mechanical efficiency, m-Pi / P is the external mechanical efficiency, iii-iihtiv-

P _r Pi) is the internal efficiency, and P, _Pr , Pi are the shaft power, the internal mechanical loss power, and the internal power, respectively. The internal friction reduction scheme of the present invention reduces the friction loss of the disk by an order of magnitude, so the internal mechanical loss power P _r is also proportionally reduced to (Zi / Z ₂ ) P _r , and the internal power Pi also becomes Pi-(Z ! / Z ₂ ) P _r , so the internal mechanical efficiency, internal efficiency and overall efficiency have been improved accordingly.

 Under the premise that the external mechanical efficiency, hydraulic efficiency, and volumetric efficiency are unknown and it is assumed that they remain unchanged, after a little more derivation and calculation, the efficiency gain data of the internal friction reduction technology under the assumed operating conditions can be obtained as shown in Table 8. 9. Because the efficiency loss caused by the friction loss of the disc is quite different, it must be discussed and calculated separately. Tables 8 and 9 show the efficiency improvement data of friction reduction in small or low specific speed centrifugal pumps and large centrifugal pumps, respectively.

 Data sheet for efficiency improvement after friction reduction in small, low specific speed centrifugal pumps

 80% 70% 60% 50% 40%

 Effective 10% 70% 60% 50% 40% 30%

 Total efficiency 7 & 0% 68.0% 575% 47.8% 37.7%

 82% total efficiency increase + ao% + ao% + 75% + 7.8% + 7.7%

 Total efficiency 785% 6 & 9% 58.8% 48.8% 387%

From 90% total efficiency increase + & 9% + 83% + & 8% + & 8% + 8.7% Referring to Table 8, the first and second rows in the table assume friction reduction conditions. Wheel friction causes a 10% reduction in nominal efficiency under several efficiency possibilities. The last 4 rows show the friction reduction efficiency data when the friction loss ratio is 0.18 and 0.10. In the prior art, the friction loss of the disk of a small, low specific speed centrifugal pump is relatively serious, which may cause a nominal efficiency drop of about 10%. When running at a low flow rate, the efficiency drop may reach more than 20%. After the internal friction reduction scheme is adopted, the ratio of the inflation diameter of such pumps can reach 96%-98%. According to Table 7, the friction loss ratio of the disc will be between 0.18 and 0.1. As shown in Table 8, the nominal efficiency will increase by 8-9%. When the efficiency is high, the increase will be large, and when the efficiency is low, the increase will be slightly reduced.

 Data sheet for efficiency improvement after reducing friction in large and medium-sized centrifugal pumps

 Referring to Table 9, the first and second rows in the table assume friction reduction conditions. Under several efficiency possibilities, wheel friction causes a nominal efficiency reduction of 6%. The last 4 rows show friction reduction efficiency data when the friction loss ratio is 0.10 and 0.05. In the prior art, the disc friction loss of large and medium-sized centrifugal pumps is very large, but the efficiency loss caused can generally be controlled at 6% or less. When running at low flow rates, the efficiency drop will exceed this value. After the internal friction reduction scheme is adopted, the ratio of the inflation diameter of such pumps can reach 98% to 99%. According to Table 7, the friction loss ratio of the disc will be between 0.10 and 0.05. As shown in Table 9, its nominal efficiency will increase by 5.3% to 5.7%. The increase is positively related to efficiency, but the difference is small.

Internal friction reduction technology also has obvious advantages in balancing and stabilizing axial forces. For the internal friction reduction scheme of the closed impeller, due to the pressure difference of the gas phase centrifugal force field Very small, so the end cavities on both sides can be considered to have the same specific pressure, and the pressure difference is only in the size of the pressure bearing area. Considering that the impeller inlet diameter is usually about 1/3 of the impeller diameter, excluding the shaft cross-sectional area, the pressure-bearing area of the end cavity on the suction side of the pump will be only 5% to 9% smaller than the pressure side, plus the fluid flow on the suction side. The compensation of the steering dynamic reaction force, during normal operation, the difference between the axial forces on both sides of the impeller may drop to the level of 2 to 5% of the pressure on the pressure side. Moreover, this pressure difference will only change slightly with the change of the flow rate, without the disturbance of other uncertain factors, so it is easy to balance and control. For a semi-open impeller, after the friction is reduced in the pressure side, the axial thrust generated will no longer be affected by the gap leakage flow and will become constant. Although there are still many related changes on the suction side, the balance effect is worse. However, the stability of the axial force will be significantly better than the non-aerated impeller.

 The internal friction reduction method and its device are important components of the invention design for comprehensively improving the efficiency of the centrifugal pump of the present invention, which can be used in combination with other invention features or can be implemented separately, so it has the property of an independent invention.

 When this design is applied to the prior art centrifugal pump, it is equivalent to overcoming the third largest bottleneck that restricts efficiency. As shown in Tables 8 and 9, it can improve the efficiency by 5-9 percentage points, which is already considerable. Benefits, but implementation costs are low. Since the prior art centrifugal pumps do not have such a structural plan, many of the end cavities are large open openings, which makes it difficult to inflate and maintain the stability of the airbag. Only when the opening of the end cavity is reduced to a small gap can the prior art centrifugal pump be conveniently installed with an inflation device, which will be provided in the description of the embodiment of the present invention. When the internal friction reduction technology is applied to centrifugal pumps with high potential ratios, the problems of the first major bottleneck diversion loss and the second largest bottleneck impeller loss that restrict efficiency have been resolved, and the hydraulic efficiency and total efficiency basis of centrifugal pumps have been greatly improved. . As a result, the friction loss of the disk became the main factor restricting the efficiency and rose to the first bottleneck. At this time, the use of inflatable friction reduction technology has increased the efficiency slightly.

The specific design of antifriction in the present invention includes the choice of gas type and the reduction of end cavity pressure And the inflation diameter ratio increase method, the circulation drive and the flow adjustment solution. The choice of gas type should meet the requirements of no harmful physical and chemical reactions with the liquid being pumped and easy to obtain. For example, when the liquid being pumped is water, using air is the simplest choice. The existing technology already uses air to Application precedents for regulating specific work or improving starting and stopping transition characteristics. The end-cavity pressure and the level of liquid level disturbance are of course as low as possible. They are related to the geometric design of the impeller outlet static pressure, the outlet flow velocity and the exit part, and also to the end-cavity clearance. The design should be optimized through special experiments. The present invention A specific scenario related to this will be provided. As for the cycle driving method, it should be comprehensively considered according to economic factors such as end-cavity pressure, gas source pressure, and cost. It can be selected from the schemes of pressure gas cylinders, gas pumps, and gas-liquid two-phase flow pumps. Although gas circulation is necessary, there is no strict quantity requirement for the circulating flow rate, and it can be adjusted to the minimum value in the premise of not affecting the friction reduction effect. The circulation of the pressure liquid is also necessary. This is the need to cool the shaft seal and create a pressure difference between the leakage gaps, or it may be the energy source driving the two-phase flow. The circulation flow rate should be determined based on the maximum flow rate of all these needs. In fact, in terms of cyclic drive for internal friction reduction and its flow control, there are many options available.

A specific solution for internal friction reduction in the present invention is: a jet device for inflating a friction reducing end cavity is provided, and the driving pressure liquid of the jet device is divided by an outlet of a pump, and its ejection port is connected to an air source or through the atmosphere through a regulating valve, and its outlet is The output pressure is slightly higher than that of the end-cavity gas-liquid two-phase flow, which is connected to the anti-friction end-cavity from the paraxial axis of the stationary wall. The two-phase flow is separated in the cavity. The gas is confined in the cavity by the centripetal buoyancy of the centrifugal force field. Excess gas is expelled from the wheel gap into the deflector. When reducing the friction in the front chamber, the gap through the suction chamber is changed to a gas-blocking gap to prevent gas from escaping, and a gas-blocking V-ring groove that separates and splits the two-phase flow is installed at the inlet, or a two-phase flow sealed and lubricated organic is added. A material retaining ring or another pressure liquid is used to directly seal the gap. The end face of the disc is either roughened to increase the peripheral speed of the medium. The cavity wall generatrix at the exit of the impeller is also designed to have an ejection decompression effect, so that the ejection effect of the outlet liquid flow reduces the end cavity pressure to Outlet static pressure Below the force, the driving liquid flow rate and the induced gas flow rate are adjusted in real time, and the end-cavity gas-liquid interface is stabilized near the wheel edge.

 The ejector works on the principle of momentum exchange. Although its efficiency is not high, its cost is very low, and its volume is very small. It has been widely used in the pressure swing operation of small flow fluids. Since the pressurized liquid and its pre- and post-processes are readily available in the main equipment, it is a perfect design to use a jet to drive the gas to achieve the internal friction reduction of the centrifugal pump. Practice has shown that when the pressure of the driving liquid flow is higher than the target pressure by more than 0.05 MPa, the ejector has enough ejection kinetic energy to produce a significant ejection boosting effect. The parameter that tests the jet effect and affects the efficiency is the working head ratio ε, which is determined by the formula (8).

Wherein ^, Ρ _2, Ρ ₄ are the input fluid pressure entrained flow pressure and output pressure of the body, according to the connection scheme of the centrifugal pump outlet pressure, respectively, the gas pressure source and target-side pressure chamber. The characteristics of the ejector are: The lower the working head ratio, the smaller the required fluid flow rate. It can be known from formula (8) that reducing P ₄ and increasing P ₂ within the limits of Pi are measures to reduce the head ratio of the ejector, and the case where (P ₄ -P ₂ ) is negative is considered and included. Therefore, reducing P ₄ and increasing P ₂ within the limits can reduce the flow of the pressure liquid.

When the cavity wall busbar at the exit of the impeller is designed to have the effect of ejecting decompression, the impeller and its containing cavity become a large static pressure regulating ejector, and its ejecting effect can reduce the pressure at the end of the friction chamber P ₄ Reduced to below the impeller outlet static pressure, which helps the ejector reduce the target pressure. This is a static pressure adjustment that does not consume power. If the solution is to be implemented in the final stage of a single-stage pump and a multi-stage pump, reducing the head ratio by using the jet effect of the impeller is an important measure to reduce the end cavity pressure and the feedback pressure liquid flow. There is another important effect of reducing the end cavity pressure, which is to reduce the axial thrust, because this thrust is proportional to the end cavity pressure.

When the end cavity is inflated, it is important to maintain the stability of the liquid level (radial coordinate of the gas-liquid interface) and the pressure, which needs to eliminate or reduce the influence of various disturbance factors. Above party The case requires that the end cavity side of the wheel disc should have a rough surface, and its role is equivalent to the installation of countless tiny auxiliary blades on the outside of the impeller, which can increase the driving force for the fluid in the cavity and enable the The liquid has a higher peripheral velocity, and the two-phase flow into the input cavity will also increase the peripheral surface velocity. If the liquid in the two-phase flow falls directly into the gas-liquid interface without such acceleration, it will cause a velocity shock and affect the stability of the liquid level. In the above solution, the flow channel area and direction of the two-phase flow entering the end cavity are slightly controlled, and a two-phase flow velocity sprayed to the end face of the impeller can be obtained. The ejected liquid will be split into small liquid beads attached to the rough end face of the impeller, and the radial velocity will be increased by the viscous force in the radial relative movement. Liquid bead acceleration can reduce the amplitude of the liquid level disturbance and prevent gas from escaping, thereby reducing the flow of gas and pressure liquid, and achieving the purpose of improving volumetric efficiency.

 There is an annular leakage gap between the front end cavity of the closed impeller and the suction chamber. When a two-phase flow is passed from the paraxial portion, the gap becomes a parallel leakage branch, so it must be changed to a gas blocking gap to prevent gas escape. . Adding a gas-blocking V-ring groove that separates and splits the two-phase flow at the inlet of the two-phase flow, or installs a two-phase flow to seal and lubricate the organic material retaining ring, or another pressure liquid directly seals the gap, which can improve the inlet. Absolute pressure to prevent gas from escaping.

 A multi-stage pump assembled by a closed or semi-open impeller requires a separate ejector and its two-phase flow cycle for each stage to maintain the gas phase pressure in its end cavity. The driving pressure liquid can be obtained from the subsequent stage. Can increase the working head ratio of the ejector and reduce the flow of pressure liquid. Of course, this method cannot be used in the final stage.

 The semi-open impeller has no front cover and front cavity. The two-phase flow driven ejector only needs to drive one cycle of the rear cavity, and its driving flow can be reduced by more than half.

 The internal friction reducing two-phase gas circulation can also be driven by the gear pump pressurized liquid flow, or the gear pump re-pressurizes the output liquid of the centrifugal pump. The ejector of this type of drive has a high working head ratio and a small pressure liquid flow. .

Inflatable antifriction scheme can also not use the two-phase flow circulation method, but use pressure tank gas to reduce the pressure of the pressure reducing valve and the regulating valve to adjust the flow rate, and then directly rest from the antifriction end cavity Wall injection, the well shunts a small flow of liquid from the pump outlet and directly injects into the mechanical seal cavity and the front-end cavity leakage gap to cool and block the leakage gap, respectively, or after the pressure canned gas is depressurized by the pressure reducing valve and the flow is adjusted by the regulating valve The airflow is directly injected into the return pipe leading from the outlet of the pump to form a two-phase flow, which is respectively connected to the rear wall static wall surface and the front cavity static wall surface paraxial gas-blocking gap, respectively, and can also achieve the same effect. At this time, the ejector is no longer needed to generate two-phase flow, and the gas and liquid flow rates are adjusted separately. The liquid flow adjustment valve is connected in series in the pump outlet shunt pipe.

 The introduction of pressure tank gas or other pressure source gas can simplify the driving device of the internal friction reduction scheme, and its work is more stable and reliable. For most water pumps, canned compressed air can be used as the pressure gas source. For pumps that pump flammable liquids, cheap canned nitrogen can be used as the pressure gas source. The gas can also be installed in the output pipeline. The gas-liquid separation cavity is cleared by a single gravity separation. A well-designed internal friction reduction device can achieve a high inflation diameter ratio, and a few percentage points of efficiency improvement during operation can save a lot of energy costs, but the consumption of gas sources is very small, and its cost is insignificant. For example, for a shaft work as

When the efficiency of a medium-sized water pump of about 100KW is increased from 60% to 64%, the electric power saved is 10.42KW, which can save 250 degrees of electricity every day and night. The gas flow consumed by the drive cycle of the antifriction device only needs about 0.1 standard state liters / second to meet the requirements. Converted into energy efficiency, when the end cavity pressure is 0.4 MPa, the compressed air flow rate of 0.1 standard state liters / second is equivalent to an effective power consumption of 0.1xl01.3xln (400 / 101.3) = 13.9W, considering 30% air source compression The efficiency is only equivalent to a power consumption of about 46.4W, and the energy input-output benefit ratio is 225 times.

The reason why the compressed air flow consumed by the driving cycle is so small is because the mechanical mechanism of the end-cavity centrifugal force field for gas-liquid separation and gas preservation is super stable and non-consumption, except for the small random disturbance of the peak value of the gas-liquid interface. By nature, maintaining a small circulating flow is only for completing the initialization process and further increasing the inner diameter of the disturbed liquid seal ring during steady state operation. In fact, the design of circulating flow is not Too strict quantity requirements can even be on the order of magnitude of change. When the design circulation flow is small, the impact is only the lengthening of the initialization process time and the depth of the liquid seal ring may increase. The former effect can be ignored in the steady-state time length and does not affect the steady-state operating conditions. The latter has a probability effect. Steady-state conditions, but the liquid seal depth is limited by the small amplitude of the disturbance peak, which has little effect.

 The solution to the third object of the present invention is: The deflector is a centripetal guide wheel, and the liquid flow flows from the outer edge to the confluent outflow of the paraxial annular cavity. The flow channel is inwardly swirling, the radius of curvature gradually decreases and the cross-sectional area gradually expands. The sum of the cross-sectional areas at the inlet or outlet is equal to the design flow rate divided by the design flow velocity at the inlet and outlet. The flow channels are distributed symmetrically in rotation, and are divided into a transfer section and a boosting section. The transfer section is connected to the exit of the impeller flow path, and the cross-sectional area of the boosting section is continuously expanded, converging in the central ring cavity, and outputted axially after turning 90 degrees. The runners are separated by guide vanes, which are installed or integrated on the base plate to form a semi-open structure or a cover plate to form a closed structure. The center of the base plate is provided with a shaft sleeve which is dynamically matched with the rotating shaft.

 In the prior art, a worm wheel or a guide wheel and a guide ring are used to directly guide and pressurize the outer ring space. When a volute is used, the confluence with different speed distributions and the supercharging process share a non-completely constrained flow channel. Conflicts in the speed distribution of supercharging and supercharging, and the large impeller openings cause vortex abduction of the volute, which can lead to instability of the flow field and local excitation loss, so the volute has the lowest flow efficiency. When a guide wheel and a guide ring are used, the confluence and pressure increase processes still coexist in the large open communication space, and the two are in an under-restricted state that is connected in parallel. The large opening leads to the relative vortex and other unstable flow state abduction of the impeller, and its return area even penetrates deep inside the guide ring, and the trailing edge vortex interferes with the interface flow field distribution. The under-constrained state on both sides of the impeller leads to local excitation loss, and the flow efficiency is not high.

When cascading overcurrent is required, the steering mechanism of the prior art guide wheel, which is mainly composed of the axial surface speed component, needs to eliminate most of the loop volume, and then quickly turn 180 degrees to enter the anti-missile process. The speed of the anti-missile flow channel changes greatly during the loop removal and anti-pre-rotation. The cross-sectional area of the anti-flow channel decreases with decreasing radius, and the anti-missile acceleration process forms a centripetal pressure reduction distribution. Oversubtraction The unreasonable phenomenon of accelerating after the speed increases the diversion load and is one of the factors that reduce the efficiency

 Both types of traditional deflectors do not have adaptability to changing working conditions. When the flow rate decreases, the local excitation phenomenon is aggravated, and the diversion efficiency is greatly reduced. In terms of space utilization, the deflectors of traditional single-stage or multi-stage pumps are looped outside the impeller, which increases the radial size of the pump and is therefore bulky. Among them, the anti-deflection wheel of the multi-stage pump also occupies another adjacent axial space and increases the axial size of the pump. The voluminous worm wheel, guide wheel and anti-guide wheel waste space and increase manufacturing costs.

 The centrifugal guide wheel constructed in accordance with the principle of the whole-cycle conservative loop design of the present invention is an inward scroll type deflector. The booster flow path is fully constrained, and the sectional area is gradually increased and the center line curvature is gradually reduced according to the optimized expansion rate. Radius, small volume and high diversion efficiency. The pressure distribution characteristics of a centripetal guide wheel are exactly the opposite of traditional deflectors. The pressure of the flow channel increases monotonically with the decrease of the centerline and the radius. This guide wheel has the best matching relationship with the impeller. Its inwardly swirling flow path and its parallel arrangement with the axial direction of the impeller fundamentally change the structure and spatial position of the traditional deflector, compared with the outward scroll. Flow channel and traditional deflector with radial impeller, the guide wheel has outstanding advantages in terms of efficiency and cost.

The design of the present invention is particularly advantageous for improving efficiency for three reasons: First, the impeller flow and the diversion flow are truly geometrically connected, and thus the confluence process and the pressurization process are separated. Regardless of the impeller flow field, the diversion flow is caused by It is completely isolated in series and has complete constraints, so there is no local excitation loss. After the damage is eliminated, what remains is only controllable loss along the way. Second, the cross-sectional area expansion rate of the pressurized runner can be independently changed and optimized, and its cross-sectional shape can also be optimized, so that the loss along the way can be reduced to a very low level, and its maximum pressurization efficiency can reach 98%. Twenty-three, the complete constraint makes the flow direction independent of the flow rate, has the best adaptability to changing conditions, and its loss along the way is proportional to the third power of the flow rate, and it is in a very low loss state at small flow rates. These three efficiency enhancement mechanisms are compared with the three shortcomings of the traditional outer ring worm or guide wheel: the geometry of the combined boosting process is parallel or continuous, and the boosting flow path is incomplete Constraints, the expansion rate cannot be optimized, and the small flow loss increases sharply without adaptation to the operating conditions. The significant difference is needless to say.

 For the cascade of multi-stage pumps, the scheme of the present invention can better reflect its superiority, negating the unfavorable design of de-loop load and anti-pre-rotation and increasing power, and instead conserving the positive-pre-rotation cascade, just playing the conservative loop design Optimization potential. In addition to the three advantages mentioned above, this cascade also adds two advantages: a drastic reduction of the diversion load and the elimination of the 180-degree commutation link. Among them, the former includes the two factors of reducing the deceleration amplitude and eliminating the over-deceleration process due to the conservative output loop. As mentioned earlier, reducing the diversion load generally has the sensitivity of reducing power by a third power function. The latter means that the 180 ° axial surface component commutation link is eliminated by maintaining the peripheral velocity overcurrent. The impact is multifaceted. In addition to eliminating local power loss of the second power function type, it is more important to save The two benefits of the anti-missile process and the over-deceleration process are key improvements that affect the basic process planning and structural layout, and have a significant impact on space utilization. It should be noted that the process of reversing the axial component of the solution of the present invention is no longer non-existent, but is decentralized. It is easy to arrange a relatively small value of the axial speed increase vector over a long distance and a long time, and its spatial change rate and time change rate are small, so it does not have the geometry and mechanics of sharp turning Characteristic, its local resistance loss is thus eliminated, which is the goal benefit of conservative loop design.

 Reduction of the diversion load and its space overhead, and the supercharging process that achieves a monotonic reduction in speed without causing excessive deceleration, are all technical factors that greatly reduce the diversion process. Therefore, at the same speed, due to a number of factors, such as reduction of loss along the way, elimination of partial loss of steering, and realization of continuous deceleration and supercharging, the hydraulic loss of diversion will be greatly reduced. As mentioned above, the high potential-to-dynamic-ratio impeller of the present invention will greatly reduce the speed of inductive guidance. When combined with a centripetal guide wheel, the above-mentioned advantages of reducing losses will be more prominent. In terms of quantity, it will be expressed as a third power Functional drastic reduction.

The matching component of the present invention and the combination with the centrifugal guide wheel can be used not only for the high potential ratio centrifugal pump, but also for the tube single transformation of the prior art centrifugal pump. Diversion efficiency is applicable to formula (2).

 A more obvious advantage of the present invention is that the space overhead of the deflector will be greatly reduced due to the advantages of this structure. The combined structure of the centrifugal guide wheel and the impeller guide wheel is compact, and the effect is outstanding in reducing manufacturing costs and facilitating use. The centrifugal guide wheel is an optimized solution specially designed according to the rationality concept of the spatial planning of the centrifugal pump, especially the multi-stage pump. Compared with the prior art, this solution saves the outer ring space of the impeller and the anti-missile space in the multi-stage structure, and instead installs the guide wheel by only the portion of the anti-missile space that is equal to the diameter of the impeller. Therefore, the radial size of the centrifugal guide wheel will be equal to the impeller, which makes the diameter design of the entire pump only need to consider the impeller diameter as a reference, and its guide space utilization rate will be increased by about 2 times, that is, the space requirement will be reduced by 2 / 3. The manufacturing cost of the pump will undoubtedly be greatly reduced. If it is used in conjunction with a high potential ratio impeller, its volume will be reduced by about 50% as the diversion load is reduced to a fraction.

 These advantages of the centrifugal guide wheel have a decisive effect on the isomorphic concept of the single-stage pump and the multi-stage pump to be disclosed in the subsequent description, as well as the establishment of a modular combination method.

 An important additional design of the centrifugal guide wheel of the present invention is: The guide wheel flow passage transfer section is surrounded by the impeller exit cylinder surface, the curved surface of the front wall of the impeller cavity and the curved surface of the front bottom surface of the impeller, and the cross section is divided into the impeller cavity portion. And the guide wheel part, the two parts are assembled and joined together. The position of its composite section is periodically shifted in the direction of the guide wheel, and its cross-sectional area increases linearly with the increase of the diversion center angle. Its period is equal to the center angle of a diversion channel, and its increasing proportionality factor is equal to the ratio of the design value of the volumetric displacement of the impeller over a unit angle to the design value of the absolute velocity of the liquid flow outlet, or it is multiplied by a value greater than 1 and less than Expansion coefficient of the minimum expansion rate of the guide wheel booster runner.

The flow channel of the transfer section receives the impeller outlet liquid flow, which belongs to the high-speed flow channel. Its design has an important impact on the overall hydraulic efficiency of the centrifugal pump. Poor design may also produce water hammer chatter effect or cavitation cavitation effect, so more careful design is required. . Its general design ^ is: The length of the flow channel and the perimeter of its section should be as short as possible, because The circular section is the most ideal. When a rectangular section must be used, its aspect ratio should be as close to 1 as possible, and it is best to fillet it. When other functional sections must be used, the length of the frictional margin line should be minimized. In addition to these general principles for controlling the friction area, the rate of change of the cross-sectional area must be controlled to meet the requirements of the confluence, and the rate of change of the marginal coordinates of the cross-section must also be controlled. These are the key variables to complete the dynamic process of the overcurrent transfer. To avoid redundant descriptions, the optimization of the cross-sectional shape and the control of its marginal coordinates will be described in the design of subsequent embodiments with reference to the drawings. The objective of the above-mentioned additional design of the present invention is to meet the over-current requirements of maintaining the speed efficiently or further including continuous deceleration and boosting. Theoretically, by using the characteristics of incompressible fluid, according to Newton's equation and Bernoulli's equation, by controlling the change of the cross-sectional area of the flow channel, and simultaneously controlling the area distribution and shape of the two parts of the cross-section, it is possible to achieve a process with any mechanical characteristics. Stream transfer. Starting from the purpose of stabilizing the average flow velocity of the liquid flow, the present invention designs a potential flow process of constant pressure, constant pressure, or deceleration. The basic method is to construct the flow channel from the minimum cross-sectional area, using a periodic process method in which the continuous deformation of the confluence cross-section is accompanied by a linear increase in area and the maximum value is obtained after the two cavities are divided smoothly. The transition of the flow channel and the rationalized control of the cross-section during the central angle period corresponding to each flow channel, including the principle of satisfying the continuity condition of the flow velocity and the acceleration minimization, are based on the thinking of the momentum theorem. Among them, the peripheral speed of the confluence process is kept constant or respected as much as possible, the purpose is to avoid a large gradient distribution in the same penetrating space and cause momentum exchange loss, and local excitation is not allowed. Marginal friction loss is reduced by friction area control that minimizes the length of the margin line. The realization of the periodic process itself actually includes the periodic control of the axial component of the flow velocity. In the above design, the improvement of the flow channel's efficiency characteristics, anti-cavitation characteristics and adaptive characteristics of variable operating conditions is the focus of attention.

The flow channel scheme of the transfer section of the present invention is only embodied in the shape design of a functional curved surface on the bottom surface of the guide wheel and the outer edge of the impeller cover plate or impeller cavity cover plate, and does not form independent parts. Therefore, the manufacturing process is only equivalent to A craft shape. When involved And when the parts are mass-produced using a molding process, their implementation costs are very low.

 The design of the packaging structure matched with the centrifugal guide wheel of the present invention is as follows: (1) The central vortex and the symmetrical end caps of the divergent angle outlet tube are used for the front and rear axial packaging. The end cap is composed of a pressure-bearing cover plate with an assembly stop, a central structure integrated with the cover plate, and a connecting pipe. Its central structure includes a shaft sleeve, a central volute around the shaft sleeve, a bearing cavity and a shaft hole supported by the volute envelope structure. The pressure-bearing surface of the pressure-bearing cover is a flat surface or a rotating curved surface that matches the opening surface of the centrifugal guide wheel, and the paraxial part has an annular opening communicating with the worm. The central worm is a three-dimensional worm with a radial involute axial translation. The beginning of the worm is a tongue on the plane of the annular opening, and its end is below the tongue with increased radial and axial coordinates. The cross-sectional area of the worm is directly proportional to the center angle, and the proportionality factor is equal to the ratio of the volume displacement of the impeller over a unit angle to the average speed of the liquid flow. Taking the open circular plane as a reference, as the cross-sectional area increases linearly, the radial and axial coordinates of the centerline of the bottom of the worm gradually increase, forming a snail bottom slope, turning 360 degrees into the lower part of the tongue, and subsequently connecting Kiss. The shape of the cross section of the worm also changes with the center angle. Starting from the straight segment of the tongue, the first is the shortened semi-axis and semi-ellipse with the long axis in the plane of the opening. The involute curve with a suitable curvature change rate extends into a curved quadrilateral plus a semi-circular shape, until it enters the lower part of the tongue, and then maintains a cross-sectional area to deform into a circular cross-section to meet the pipe. When this symmetrical end cap is used as the back cover, the axial incoming flow with a circular amount enters the volute from the annular opening, converges on the vortex during rotation and translation, and rotates along the worm ranging from 0 to 360 degrees. Flow out from the kissing tube. When the front cover is used, the linear flow in the incoming tube receives the normal force of the worm and turns into a three-dimensional vortex motion. Finally, it is diverted from the annular opening into the suction chamber of the impeller. The vortex movement in the volute of the front and rear covers is low loss at the hydraulic standard speed. When the front and rear covers are rotated along the stop during assembly, the angles of the inlet and outlet pipes can be changed independently.

The symmetric end cap of the central volute branching and merging angle outlet pipe is an important supporting component of the present invention. The application of a centrifugal guide wheel with this component can make the volume of the centrifugal pump large In order to reduce the size, the function of changing the angle of the pipe can meet the unpredictable needs of users in the application. It can save space and pipes, reduce elbows and improve pipeline efficiency. This design may even make the division of the inlet and outlet angles of centrifugal pumps into different types unnecessary. Obviously, from the user's perspective, this division is inconvenient. This design goal of the present invention includes multiple versatility: it is universal in one pump, the front and rear covers; it is also universally interchangeable between single-stage and multi-stage pumps. When used in single-stage pumps, the overall advantages of this package structure are even more obvious. The multiple versatility brought by the symmetry of the end caps and the potential advantages of simplifying the model of the inlet and outlet angles can simplify the design and manufacturing process of the centrifugal pump, thereby reducing manufacturing costs.

The matching component and the combination with the centrifugal guide wheel according to the present invention can be matched with a high-potential-ratio centrifugal pump and can also be applied to the barrel modification of the prior art centrifugal pump. When it is applied to centrifugal pumps with constant potential ratio and high potential ratio, the local resistance coefficient of the volute is small, so the flow channel can be merged into the diversion flow channel to calculate the loss, and the combined diversion efficiency is uniform. Formula (2) applies. The kinetic energy loss rate ξ ₂ of the centrifugal guide wheel and its transfer section flow channel plus the symmetrical end cap will be significantly lower than that of the traditional deflector. When running under variable operating conditions, due to the characteristics of constant tangential conduction to avoid impact losses, the losses will not only increase but will decrease, so the excellent characteristic of the efficiency curve will not only decrease but increase in the small flow area. It also has special competitiveness due to its advantages of reduced size and convenient installation. Therefore, the combination of the centrifugal guide wheel and the symmetrical end cap is also an independent invention that does not depend on the high potential ratio impeller. When this technology is used in combination with inventive features such as high potential ratio impellers and internal friction reduction, the various advantageous features will complement each other.

When the centrifugal guide wheel and the symmetrical end cover are used to transform the prior art centrifugal pump, its functions and performances such as centripetal pressure increase, variable angle outlet pipe and greatly reduced volume are a prominent substantive feature, which can Produces significant progress. For example, according to preliminary estimates, only reducing the volume of the deflector and integrating the front and rear end caps and their bearing housings with the outlet pipe are symmetrical and universal, which may reduce the manufacturing cost of the centrifugal pump by more than 30%, and users can also obtain the site. Benefits such as savings, piping savings, and efficiency improvements. Obviously, this design Has a variety of technical and economic advantages.

 Based on the characteristics and advantages of the centrifugal guide wheel and the symmetric end cap, the present invention has designed a modular combined centrifugal pump method according to which many new types of centrifugal pumps can be combined.

 The invention also particularly includes a method of constructing a centrifugal pump by modular combination. This method is used not only to construct multi-stage pumps, but also to construct single-stage pumps, not only to implement one technology, but also to implement multiple technologies in combination.

 Based on the principle of centripetal supercharging of the centrifugal guide wheel, the process of the guide wheel is axially juxtaposed with the impeller speed field and its kinetic energy distribution to form the most concise periodicity of the process space. According to this periodicity, the most compact modular structure of the space concept can be established, which implies the connection mode of "fluid flow in and out of the paraxial annular mouth with ring volume." Access through the paraxial annular mouth is a necessary design The reason why the loop quantity is required is because of the mechanical considerations for minimizing the flow rate change rate of the centrifugal pump during the whole process. This kind of minimization has many benefits, including the adaptability of variable operating conditions and the improvement of inlet and outlet cavitation characteristics. The flow load is reduced, the kinetic energy of the inlet and outlet is used, and the most important is the efficient introduction and guidance after exiting from the impeller. These all belong to the optimization of the boundary conditions of the impeller flow channel. It can be seen that the conservative loop is the core problem of the optimization of the boundary conditions.

It can be known from the foregoing description that the paraxial annular opening of the symmetrical end cap and the three-dimensional volute are designed with conservative quantities, have the ability to make up the required boundary conditions, and can adapt to the above connection mode. Therefore, based on the principle, the periodicity, the module concept, and the connection mode, a method for modularly combining impellers with any technical characteristics can be designed. This method requires the external shape and size of the impeller to be standardized. Based on the standardized external shape and size of the impeller, the axial space of the centrifugal pump is expanded to design a modular assembly structure for axial connection. This structure contains all the structural elements in a hydraulic flow regime, which is actually the axial combination of the impeller, the guide wheel and its accessories. Their structure and function have periodicity of geometry and kinematics and dynamics, while the potential pressure and static pressure in the flow field are accumulated along the way. It does not have and does not require such periodicity to communicate under the principle of the communicator. The modular combination method designed in this way has nothing to do with the number of stages, so it must be a new design method that is generally suitable for single-stage and multi-stage pumps.

 The accurate description of the modular technology combination method and its combined products is: Based on the structure and characteristics of the centrifugal guide wheel, its process is axially juxtaposed with the reverse process of the kinetic energy distribution of the impeller process to form a simple spatial periodicity. An axially combined centripetal booster energizing module, which is referred to as an impeller with a ring-shaped near-axis connection, is standardized and referred to as a centripetal booster module. Based on the structure and characteristics of the symmetric end cap, it is used as a flow field boundary and end seal structure module with a ring-shaped paraxial ring mouth connection and standardized. Standardization produces two specifications series that contain child specifications. The module assembly dimensions and basic interface parameters (such as flow rate and speed) of the same parent specification are the same and have a look-up table to check interchangeability. The same seed specification under the parent specification The module assembly dimensions and all interface parameters (including parameters in the sense of functional, such as pressure resistance 2MPa, made of corrosion-resistant materials, etc.) are the same and have complete interchangeability. The two interchangeability are defined in single-stage pumps and multi-stage pumps. Various models and different built-in technologies include a large set of centrifugal pumps using high potential ratios or constant potential ratio impellers. The interchangeability domain is defined in the planning and design process and is established in the facility. According to the connection mode of "fluid flow in and out from the paraxial annular mouth with annular flow", one or a maximum of 64 are connected in series (traditional multi-stage pumps generally do not exceed 20 stages, the number of stages of the present invention can increase, but generally does not exceed 64-stage) multiple centripetal booster modules combined with 2 symmetrical end cap modules constitute a modular combination single-stage or multi-stage pump. The centripetal booster module without a shell is equipped with a mid-open type shell, and the one with a shell is a stepped structure.

In the prior art, there are no plans for different types of centrifugal pumps to use the same main functional modules, except for a few auxiliary parts such as bearings, shaft seals, and fasteners that are used across models. They also do not include different technologies and have different performance. The design of the main function modules is combined and replaced on the same centrifugal pump. As is well known, each single-stage pump in the prior art is individually designed according to the selected hydraulic model, and different hydraulic models The impellers and deflectors of the same type are not interchangeable. The impellers and deflectors with the same hydraulic model have different sizes and are not interchangeable due to different models and specifications. The prior art stage multistage pumps usually use the same energized parts or components, and they can be used to assemble different head centrifugal pumps of the same flow specification, but the models and technologies used are fixed and cannot be changed, and, Its supporting envelope structure and shafting components are still individually designed. Due to the existence of these problems, the product types of centrifugal pumps are very complicated, and the promotion of new technologies for centrifugal pumps has also been delayed by many regulatory barriers.

 The present invention expands the traditional method of assembling multi-stage pumps with the same energized components to a modular combination method, so that between the single-stage pump and the multi-stage pump, in the design of the envelope and support structure, and then in different technologies In the combinatorial algebra domain, all the modular components with advantages can be widely used. After the specificization, a new design method is formed. In the core idea of designing this method, the use of combined and modular process benefits and breaking the regulatory barriers faced by new technologies are only part of the purpose. Taking the opportunity to incorporate mathematical planning methods to improve the cost-effectiveness of centrifugal pump design is another Urgent idea. It is even more urgent because the modular combination method and its structural design may only be a partial application method. If the parameter construction module is randomly selected according to this method, it will cause a messy space occupation on the flow-pressure plane, which is unfavorable or has no great benefit to the system's system resource optimization and cost reduction. In order to benefit from the overall situation, the traditional method of directly planning the type coverage of centrifugal pump products based on social demand survey data on the flow-pressure plane should be changed, and the modular and combined structural design should be taken into account, which can greatly simplify centrifugation. The profile of the pump systematically reduces the average production cost of the pump industry. The basis for the parameter selection of the modular combination should be derived from the optimized planning specifications.

In the present invention, a centripetal booster module is used as a low-cost building block component of a centrifugal pump to establish a multi-stage assembly specification for a centrifugal pump, and a single-stage pump is also specifically included in the specification to make consumption The design and manufacture of time-consuming single-stage pumps is no longer necessary. This is the original idea. But the inventor is studying the performance equation of a multistage pump When comparing it with the high-potential ratio technology, a new technical deduction was produced. As is well known, making a multi-stage centrifugal pump is beneficial to improve efficiency, because, for the same head requirement, the multi-stage pump's stage head is inversely proportional to the number of stages, and its impeller diameter and inlet speed and the square root of the number of stages at the same speed It is inversely proportional, so the second power type guide loss of the pump will also be inversely proportional to the number of stages. Comparing the relationship between the technical characteristics of the high potential ratio impeller and the characteristics of the multi-stage pump according to the present invention, from the effect on the hydraulic loss, the design parameter feedback reduction ratio of the high potential ratio impeller gives a linear reduction function relationship, which is The adjustment parameters of the multi-stage pumps have the same direction, so they can be regarded as equivalent adjustment parameters between zones. The difference is that the latter is discrete and has a large application space, while the former is continuous but the application space is affected by limit. This equivalence reveals another important way to improve the efficiency of centrifugal pumps. The difference shows the good prospect of the complementary combination of the two approaches.

 The present invention proposes to combine the two high-efficiency approaches, and specifically integrate them in the type planning that considers the modular combination method, which can generate huge economic and social benefits. This combination is actually a combination of converting the type plan of the centrifugal pump on the flow-lift level to the level module plan, thus turning the problem into a plan of the level module on the flow-lift level. The invention designs a method for mathematically planning the wheat numbers of the level modules.

The mathematical programming method is: planning the series and specific revolutions as a binary function with flow and head as independent variables, and using the feedback reduction ratio as a continuous Tuning independent variables also participate in planning, forming an equivalent ternary programming function group ((stage number, feedback reduction ratio), specific rotation number) between the previous binary partitions. Through social surveys, an operating cost function constructed with factors such as efficiency and a manufacturing cost function constructed with factors such as structure, size, material, and process are added, and the value of the comparable unit function is added as the objective function, or the aesthetic design of the product is also increased Constraints such as additional inequalities are required, and classical mathematical programming methods or numerical algorithms are used to obtain the optimal range of the ternary programming function group that minimizes the objective function, thereby establishing a hydraulic model and a series of modular module specifications. When designing, take head and flow as the parameters, and determine the number of stages, specific revolutions and feedback reduction ratio according to the specifications to determine the optimized centrifugal pump stage module. PT / CN2004 / 001413

parameter. When the stage module adopts the backward curved impeller of the prior art, its feedback reduction ratio is defined as the ratio of the circumferential component of the relative exit speed to the line speed along the wheel, and its value is close to 0, which can be ignored.

 In the above method, due to the equivalence between the partitions of the series and the feedback reduction ratio in the efficiency adjustment, a complementary connection relationship, the level function and the specific rotation plan are provided on the same surface of equal efficiency. The resulting optimized range will be thinned accordingly. This method opens up a new thinking space for the centrifugal pump's spectrum planning and product design, which is of great significance in both theoretical research and design practice. This method has the least number of stages and specific rotations. The centrifugal pump is designed not only with high efficiency, but also with high space utilization, small size, and material savings. Therefore, it has the dual value advantages of efficiency and cost. In theory, its technical and economic indicators will be the highest.

It is important to point out that the pump efficiency and pump cost are not the final evaluation indexes. The final index should be the corresponding index of the unit, and the actual point should be the corresponding index of the actual liquid flow system. From the pump index to the final application index, there is a network topology relationship, which must be analyzed using matrix operations. For example, from the perspective of a manufacturer, if its product is a motor-centrifugal pump unit, the increase in pump efficiency as a multiplicative factor obviously increases the unit efficiency in proportion. But this is not the only result, because there is another design function that affects the interests of manufacturers, that is, the relationship between pump efficiency-motor power-motor cost-unit cost. Therefore, the unit cost consisting of the sum of the cost of the pump, the cost of the motor, and the cost of the connection between the two will be affected by both the pump efficiency and the pump cost. Considering the proportion of the motor cost in the cost of the centrifugal pump unit, the significant increase in the efficiency of the pump on the unit cost is not small. It can be seen that in the technical and economic evaluation, the pump index is just a piece of vector. In the coefficient matrix for solving the target index related to actual economic benefits, the influence of cross-acting factors exists and should be taken into account. For end users, the impact will be on their total cost of ownership in terms of both purchase costs and operating costs. For manufacturers, a design that simultaneously benefits both performance and cost will undoubtedly increase the product ’s performance. Cost-effective and increase its market competitiveness. Based on this consideration, the objective function in the above planning method should be transformed, and the method should be revised to: Obtain an association matrix function through survey statistics, convert the pump efficiency and pump cost into unit efficiency and unit cost, and construct the unit operation Costs and manufacturing costs, and the values of the comparable unit functions are added as an objective function for planning to optimize module parameters and product design.

 The invention pushes the idea of multi-level structure and modular combination into consideration a process premise, that is, new technologies provided by precision molding technology and other modern manufacturing technologies, such as powder metallurgy, pressure precision casting, compression molding and injection molding. Compared with traditional processes, these processes can increase the space complexity of the structure, and at the same time can greatly improve production efficiency and reduce production costs. For example, the semi-open impeller, closed impeller cover, guide wheel with casing and impeller cavity, impeller cavity cover and other parts in the step-type impeller combination module are all open or semi-open workpieces, which have a general ability to Adopting the advantages of the two-clamp molding process, the new process can be used for manufacturing. The series of hydromechanical feature designs of the present invention may increase the complexity of the mold and the cost of the mold, but basically do not increase the processing cost. Considering that the cost value of the mold is actually low in mass production, it is not only the inventor's desire to implement the series of technologies of the present invention with a new process, but also an implicit basic assumption. If traditional processes are used, some designs of the present invention will be difficult to implement or the cost will increase. This will not affect the practicality of the present invention, because society urgently needs cheap and efficient centrifugal pump products. The combination of the design of the present invention and appropriate advanced technology can meet this demand.

 By adopting the modular combination method of the present invention, the aforementioned invention characteristics of various types of products can be combined into a module, and the inventive features of some of these products can also be combined with existing technologies, which will form many new types of centrifugal pumps. Modification of existing technology product designs using only a modular combination approach can also produce a variety of positive effects. The detailed design of these combinations and the description of the combination effects, as well as the application methods of the combination methods, will be included or embodied in the description of the embodiments with reference to the drawings.

BRIEF DESCRIPTION OF THE DRAWINGS The centrifugal pump and its technical combination method and examples according to the present invention will be further described in detail below with reference to the accompanying drawings.

 Figure 1 is a schematic diagram of a semi-open type high potential ratio impeller.

 Figure 2 is a schematic diagram of a closed high potential ratio impeller.

 FIG. 3 is a schematic diagram of a structure of a zigzag disc semi-open high potential ratio impeller.

 FIG. 4 is a schematic diagram of a structure of a closed-type high potential ratio impeller with a sawtooth-shaped disc. .

 Fig. 5 is a schematic structural view of an axial inflow pre-spinner.

 Fig. 6 is a schematic structural diagram of a radial inflow pre-spinner.

 Figure 7 is a schematic diagram of a semi-open type high potential ratio impeller with a pre-spinner.

 Figure 8 is a schematic diagram of a semi-open type high potential ratio impeller with a uniform speed bifurcation. Figure 9 is a schematic diagram of a semi-open high potential ratio impeller structure with a uniform speed bifurcation and a pre-rotator.

 Figure 10 is a schematic diagram of the structure of a high potential ratio cantilever pump with a uniform speed bifurcation and a pre-rotator.

 Fig. 11 is a schematic view showing the structure of a two-phase flow cooling shaft seal flow channel with antifriction driving inside the shaft end.

 Fig. 12 is a schematic diagram of the structure of a V-shaped groove air-blocking gap for driving a two-phase flow in a front-end cavity to reduce friction.

 Figure 13 is a schematic diagram of a semi-open impeller cantilever pump inflatable drive device and its connection.

 Fig. 14 is a schematic diagram of a closed impeller cantilever pump inflatable driving device and its connection. FIG. 15 is a schematic structural diagram of a centripetal guide wheel.

 Figure 16 is a schematic diagram of the flow path structure of the transfer section of a combination of a high potential ratio impeller cavity and a centripetal guide wheel.

 Figure 17 is a schematic diagram of the structure of a closed impeller super friction reduction and guide wheel control transfer section.

Figure 18 is a schematic illustration of the structure of a symmetric end cap of a centrifugal sub-confluence rheological outlet pipe Illustration.

 Figure 19 is a schematic diagram of a stepped modular structure of a combination of a semi-open impeller and a centrifugal guide wheel.

 Fig. 20 is a schematic diagram of a stepped modular structure of a combination of a closed impeller and a centrifugal guide wheel.

 Fig. 21 is a schematic diagram of a stepped modular structure of a friction-reducing closed impeller and a centrifugal guide wheel.

 Figure 22 is a schematic diagram of the structure of a stepped module with a combination of a semi-open high potential ratio impeller and a centrifugal guide wheel.

 Figure 23 is a schematic diagram of the structure of a stepped module with a closed high potential ratio impeller and a centripetal guide wheel.

 Figure 24 is a schematic diagram of the structure of a stepped module of a pre-spin closed high potential ratio impeller and a centrifugal guide wheel.

 Figure 25 is a schematic diagram of the structure of a stepped module with a friction-reducing closed high potential ratio impeller and a centrifugal guide wheel.

 Fig. 26 is a schematic diagram of a stepped module structure of a combination of anti-friction pre-spinning high potential ratio impeller and centripetal guide wheel.

 Figure 27 is a schematic diagram of the structure of a stepped module with a combination of a super-friction reducing pre-spinning high potential ratio impeller and a centripetal guide wheel.

 Figure 28 is a schematic diagram of the structure of a semi-open wheel centrifugal booster centrifugal pump with a symmetrical cover and variable angle outlet pipe.

 Figure 29 is a schematic diagram of the structure of a symmetric cover variable angle outlet tube closed wheel centrifugal booster centrifugal pump.

 Figure 30 is a schematic diagram of the structure of a symmetric cover variable angle outlet tube anti-friction closed-wheel centrifugal booster centrifugal pump.

FIG. 31 is a schematic structural diagram of a semi-open centrifugal booster centrifugal pump with a high potential ratio of a symmetric cover variable angle outlet pipe. Figure 32 is a schematic diagram of the structure of a closed-wheel centrifugal booster centrifugal pump with a high potential ratio of a symmetrical cover variable angle outlet pipe.

 Figure 33 is a schematic diagram of the structure of a closed-wheel centrifugal centrifugal centrifugal pump with a pre-spinning high-potential ratio and a closed-wheel centrifugal pump.

 Figure 34 is a schematic diagram of the structure of a closed-type centrifugal booster centrifugal centrifugal pump with a symmetric cover and variable angle outlet tube to reduce friction and high potential.

 Figure 35 is a schematic structural diagram of a symmetric cover variable angle outlet tube anti-friction pre-spinning high potential ratio closed wheel centrifugal booster centrifugal pump.

 Fig. 36 is a schematic structural diagram of a closed-type centrifugal booster centrifugal pump with a superimposed friction reduction pre-spinning high-potential ratio and a closed-wheel centrifugal pump.

 Fig. 37 is a structural schematic diagram of a semi-open type wheel centripetal booster multistage centrifugal pump with a symmetrical cover and variable angle outlet pipe.

 Figure 38 is a schematic diagram of the structure of a symmetric cover variable angle outlet tube closed wheel centrifugal booster multistage centrifugal pump.

 Figure 39 is a schematic structural diagram of a symmetric cover variable angle outlet tube anti-friction closed-wheel centrifugal booster multistage centrifugal pump.

 Figure 40 is a schematic diagram of the structure of a semi-open centrifugal booster pump with centrifugal pump with a high potential ratio of a symmetrical cover and variable angle outlet pipe.

 Figure 41 is a schematic diagram of the structure of a multi-stage centrifugal centrifugal pump with a high-potential ratio and a closed-wheel centrifugal booster with a symmetrical cover and variable angle outlet pipe.

 Figure 42 is a schematic diagram of the structure of a symmetric cover variable angle outlet pipe pre-spinning high potential ratio closed wheel centrifugal booster multistage centrifugal pump.

 Figure 43 is a schematic diagram of the structure of a symmetric cover variable angle outlet pipe with reduced friction and high potential ratio closed wheel centripetal booster multistage centrifugal pump.

 Fig. 44 is a structure diagram of a multi-stage centrifugal pump with a symmetric cover and variable angle outlet pipe for reducing friction and pre-spinning with high potential ratio, and a closed-wheel centrifugal booster.

Figure 45 is a symmetric cover variable angle outlet tube super friction reduction pre-rotation high potential ratio closed wheel centripetal Structure schematic of booster multistage centrifugal pump.

 Fig. 46 is a schematic structural diagram of a centrifugal multi-stage centrifugal pump with a centrifugal booster pump with a pre-spinning and high potential ratio of a symmetrical cover variable angle outlet pipe.

detailed description

 Figure 1, Figure 2, Figure 3, and Figure 4 show the four main types of high potential ratio impellers and their structures, respectively.

 Referring to FIG. 1, a structure of a semi-open high-potential-ratio impeller is shown. Among them, 1 is the impeller disc, 2 is the impeller shaft hole, 3 is the impeller shaft sleeve, 4 is the suction chamber, 5 is the blade, 6 is the middle of the flow channel of the blade groove, 7 is the flow channel inlet, and 8 is the flow channel outlet.

 The half-open high-potential-ratio impeller is a disc-shaped part and is manufactured in one piece using a molding process. It has a shaft hole 2 and a shaft sleeve 3 in the center for assembly with the shaft (keyway can be set). The outer surface of the sleeve is a ring-shaped suction chamber 4. The bottom surface of the sleeve is a rotating curved surface that continuously turns the liquid flow. When the pre-spinner is installed, the surface will be replaced by the pre-spinner wheel surface with the same shape. The blade 5 is L-shaped with a radial trend in the front and middle, and a tail curved in the tangential direction, and the outside of the tail is a smooth involute cylinder or groove. Six to twelve identical L-shaped blades are evenly distributed on the wheel disc, and evenly spaced flute channels are formed between the blades. The inlet 7 and the middle 6 of the flow channel are relatively wide, and the cross-sectional area gradually decreases and turns before reaching the outlet 8. The fluid is accelerated and changed direction, and finally flows out of the impeller at a relatively large speed and a small outlet angle. As a result, the absolute speed of exports has been greatly reduced, and the momentum ratio has increased significantly. After the liquid flow exits, it is constrained by the inner wall or grooved surface wall effect. In the exit interval, a uniformly distributed flow line is formed with a uniform radial component and tangential component. The ratio does not change with the center angle. This is equivalent to the effect of a continuous opening, but there is no backflow at all.

The exit area of each impeller flow channel is proportional to the design flow, and is inversely proportional to the design exit relative speed and the number of blades. The design relative speed is equal to the product of the peripheral speed at the exit of the impeller and the feedback reduction ratio parameter K. The design parameters determined by this method are in good agreement with the measured values, because there is no backflow disturbance. 01413

Determine the parameter K according to the foregoing discussion. The centrifugal pump will have good regulation characteristics, efficiency characteristics and anti-cavitation characteristics.

 Referring to Figure 2, a structure of a closed high potential ratio impeller is shown. Among them, 9 is the rivet for fixing the cover plate, and 10 is the front cover of the impeller. On the basis of the semi-open type high potential ratio impeller, a front cover 10 is added to close the impeller to form a closed high potential ratio impeller. Closed high potential ratio The impeller flow channel is more curved and the exit is narrow. It is difficult to manufacture by traditional casting process. The semi-open impeller chassis riveting front cover is simple and easy. This method requires that 2 to 3 rivet holes are opened in the wide elbow of each L-shaped blade, and the front cover plate and the semi-open type roulette part are riveted with a countersunk head or a flat head rivet 9. The front cover is manufactured by a compression molding process, and has an inner surface that closely fits the half-open wheel disc and an outer surface of a rotating curved surface that meets the requirements of accuracy. In order to ensure dynamic balance, the rivet holes in the same relative position should be on a concentric circle and at the blade midline position. The rivet hole can be changed to a screw hole, and the front cover is fastened with a lock screw. The front cover can also be connected using spot welding. The technical characteristics and design requirements of the semi-open chassis of the closed impeller are the same as those of the semi-open impeller, and the use characteristics of the two are also basically the same. The advantage of the closed impeller is that its blade groove flow path is free of external friction and turbulent interference of the open surface, so it is closer to the theoretical characteristics. In addition, after the closed impeller is installed in the impeller cavity, two end cavities that are closed to the impeller flow channel can be formed. After using the double-end cavity anti-friction technology, most of the disc friction will be eliminated. Internal mechanical efficiency.

 Referring to FIG. 3, a structure of a zigzag disc half-open high potential ratio impeller is shown. 11 is the contour line of the teeth of the wheel. By removing the part of the blade of the semi-open type high potential impeller circular disc, the disc becomes zigzag. The line segment 11 in the figure is a tooth-shaped contour line segment from the tip of the blade to the outer edge of the subsequent blade, which is a straight line on the exit normal plane. The chassis at this line segment is thinned to form a sharp tooth tip to produce a rectifying effect. In order to maintain the strength of the tooth tip and blade tip, this impeller requires high-strength materials.

The zigzag wheel disc has higher potential than the impeller, which has higher efficiency. In general, leaves The axial restraint function outside the wheel outlet has been replaced by the impeller cavity. The curved triangular small pieces on the edge of the wheel disc become a loss due to friction on both sides, and the outside friction with the end cavity medium at the implication speed, which consumes the impeller ratio. Work, its inner side rubs against the liquid flow at the relative velocity of the outlet, and consumes the specific energy of the liquid flow. After removing these small pieces, there is only friction between the liquid flow and the cavity wall at the corresponding position, and only the specific energy of the liquid flow is consumed, and the absolute speed of the liquid flow is also less than the implication speed, which can obviously improve the efficiency.

, The analysis of the amplitude of the efficiency increase of the zigzag wheel is as follows: According to the speed triangle, the friction speed after removing the small triangle is one of its short sides, and the speed of the original two-sided friction is the long side of the involved speed and the short side of the relative speed. Is the sum of both losses. Considering that the loss is proportional to the square of the speed and the condition of β ₂ «0, the loss reduction rate per unit area is ((1 ² + 0.5 ² )-(1-0.5) ² ) / (1 ² + 0.5 ² ) = 80%. Considering that the differential of the loss moment is proportional to the product of the frictional force per unit area, the length of the circumference, and the radius, the integration is a fifth-order function of the radius, so the energy saving benefit should be considerable. For example, if the radial width of the outlet is 10% of the radius, the reduction ratio of the friction loss on the area of these small pieces relative to the friction loss reduction of the rear end cavity of the disc is

Ο.δίο, ¹ Γ ³ 2π (10r— 9) dr) / ίο ¹ r ³ 27rrdr = 0.8 (0.8190-0.7738) /0.2 = 18.1%

It is also set that the friction between the back-end cavity and the disk reduces the pump efficiency by about 3 to 5%, and the pump efficiency is increased by about 0.5 to 1%.

 The design goal of the zigzag wheel is to reduce the friction of the wheel, which can produce the above-mentioned benefits when there is no internal friction reduction device. Because the effect of the internal friction reducing device is more significant, it is not suitable to design the impeller of such a disk when the device is installed, because it will reduce the inflation diameter ratio and interfere with the operation of the device, and the gain will be lost.

Referring to FIG. 4, a structure of a zigzag disc closed high potential ratio impeller is shown. After removing the part of the blade of the semi-open circular wheel disc of the closed high potential ratio impeller and its cover, the entire blade stem becomes zigzag. The line segment 12 in the figure is a blade tip The tooth profile contour line to the outer edge of the latter blade is a Nai line on the exit normal plane. Its tooth-sharpened rectification and corresponding strength requirements are the same as those of the semi-open impeller shown in FIG. 3.

 The anti-friction effect of the zigzag disc closed high potential ratio impeller is double-sided. According to the analysis of the aperture ratio of the above example, the friction loss of the rear wheel and the front cover can be reduced by 18.1%, so the pump efficiency can be approximately Increase by 1 ~ 2%. For the same reason, this impeller is limited to use in centrifugal pumps without internal friction reduction.

 Referring to Figure 5, a structure of an axial inflow pre-spinner is shown. Among them, 13 is the margin of the impeller suction chamber, 14 is the lower rim, 15 is the upper rim, 16 is the elastic sail blade, 17 is the axial projection of the lower rim, and 18 is the blade near the bottom circle of the lower rim. Point 19 is the point near the bottom circle of the upper wheel, and 20 is the rigid rib.

 The axial inflow pre-spinner is assembled by two rims and several elastic sail blades. The rims 14 and 15 are sleeved on the impeller shaft sleeve, which can rotate independently, and their surfaces are kissed with each other to form a rotating curved surface that continuously turns the liquid flow. The number of sail-type blades is less than the number of impeller blades or a submultiple of the blades, so that the confinement of the flow channel is not too low and the friction area ratio is not too large. The sail blade 16 forms a curved triangle, and its straight edge is suspended on the rigid rib 20, and the rib is fixed radially at the entrance of the impeller blade or the front cover. Two points 18 and 19 on the curved side of the sail blade, which are close to the bottom circle below the two rims, are respectively fixed on the two circumferences, forming a point of pulling force opposite to the rigid rib.

The blade tail tip 18 has a certain bending strength, and has a directivity when fixed, so as to substantially keep the direction of the exit of the pre-spinner channel consistent with the direction of the inlet of the impeller channel. During operation, the impeller rotates the blade and the rim together through the rigid ribs, and the elastic sail-shaped blade adaptively changes into a propeller shape, and maintains a state tangent to the streamline from the entrance to the exit. This is the blade's tensile stress and bending stress. And fluid flow reaction force and centrifugal force automatically balance the result. In steady-state operation, the free-curve and free-straight edges of the elastic sail blade will change into a space curve, and the entire blade and its marginal line will remain stationary relative to the abutment of the wheel and the impeller. In various dynamic processes such as operating conditions, the blades The shape and the stress of each part will be adaptively changed, and the relative position with the surroundings will also be changed to balance with the fluctuation of the inertial dynamic reaction force of the liquid flow. After the balance, the tangent to the streamline will be maintained, and the two wheels will follow Adjust their angles. There is no requirement for matching accuracy, because the small gap does not affect the overall effect of the liquid flow, and the movement blocking force generated by the small squeeze will be released due to random disturbances, and the dynamic adjustment can always be accurate carry out.

 The axial inflow prespinner is installed in the suction chamber of the axial inflow centrifugal pump. Therefore, the suction chamber of the impeller must be provided with a circular cylindrical space to fit the prespinner. The center of the empty space is the impeller shaft sleeve, whose outer surface is a cylindrical surface or a stepped cylindrical surface, and is used for the rim of the sliding sleeve pre-spinner. The revolving surface of the hub will replace the revolving surface in the center of the impeller to serve as the guide flow. The load of the pre-rotator is very small, and there is no high requirement for the strength of the sail blades.

 The effect of self-adaptive deformation of the pre-rotor elastic sail blade is that the liquid flow always enters tangentially to the blade, and continuously changes the speed magnitude and direction along the course under the normal force of the blade. At the exit end, its speed direction It is always facing the impeller entrance. Therefore, impingement turbulence does not occur at the impeller inlet and the pre-spinner inlet, and the area of the impeller inlet can be used most effectively. As the flow changes, these characteristics remain the same.

 Referring to Fig. 6, the structure of a radial inflow pre-spinner is shown. Among them, 21 is the lower rim, 22 is the upper wheel, 23 is a rigid rib, 24 is an elastic sail blade, 25 is a rib bracket and a sleeve, 26 is the axial projection of the lower wheel, and 27 is the upper and lower blades. The point on the bottom surface of the wheel hub, 28 is the axial plane projection of the rigid ribs, 29 is the point on the blade close to the upper bottom surface of the upper rim, and 30 is the axis surface projection of the upper wheel hub.

The radial inflow pre-spinner is assembled by a disc-shaped rib support 25 with a shaft sleeve, a lower rim 21, an upper rim 11 and a plurality of rigid ribs 23 and a plurality of curved triangular elastic sail blades 24. The determination of the number of blades and the deformation principle of the blades are the same as those of the axial inflow pre-spinner. There are three main differences: First, the fluid flows radially from the cylindrical inlet and axially from the flat annular outlet. The curved triangular blades 04 001413

The straight edge parallel to the axis of rotation is the entrance edge, which is suspended from a rigid rib. Second, the points 27, 29 adjacent to the curved sides of the elastic sail blades and the upper and bottom surface circles of the two rims are fixed to the two rims, and the wheels provide balanced tension through these two points. Third, the wheel hubs 21 and 22 slide on the bracket shaft sleeve, and the diameter of the central hole is smaller than the outer diameter of the impeller shaft sleeve.

 In the figure, the elastic sail blade 24 is suspended on a rigid rib 28, and the rib is directly fixed on the disc bracket 25 with a sleeve, and the sleeve is statically fitted on the rotating shaft to provide driving force for the disc bracket and the rigid rib. . This structure has the advantages of component integrity and assembly independence, and its shaft sleeve and impeller shaft sleeve are axially pressed during assembly, the overall axial positioning performance of the machine is good, and the seal and rigidity of the shaft are improved as a result. .

 During operation, the disc holder 25 and the rigid rib 28 rotate in synchronization with the rotation shaft. When the circumferential component of the velocity of the liquid flow entering the pre-spinner is insufficient, the sail blades are hung backwards as shown in the figure, and the inlet angle thereof is adaptively changed with the streamline, so that the blade inlet is tangent to the streamline. The normal binding force of the blades provides the circumferential acceleration and the centripetal acceleration component of the liquid flow on the circumferential surface, so that the liquid flow enters a synchronous rotation state, and also provides an anti-radial drive to offset or partially offset the centrifugal force. Output shaft work increases fluid flow specific energy. The increase of the axial velocity component of the liquid flow is the result of the reaction force driven by the rotating curved surface. There is theoretically no function conversion. After the liquid flow enters a certain radial coordinate position, its peripheral velocity component will be continuously greater than the synchronous speed. At this time, the liquid flow will output specific energy to do work on the blade, and its force will help to balance the shearing of the tensile stress at the front of the blade. Direction component, keeping its root perpendicular to the axis of rotation, thereby maintaining a larger cross-sectional area of the flow channel. The position of the blade curve side tension balance points 27 and 29 is beneficial to maximize the angular displacement between the two rims and the impeller and between them when the load is the largest, so as to meet the needs of increased stress.

When the tangential component of the velocity of the liquid flow entering the pre-rotator exceeds the implication speed, the sail blade 24 will be stretched forward, contrary to the bending direction shown in the figure. The liquid flow output ratio can do work to the blade, the pre-rotator Enter the working state of the turbine. The mechanical analysis at this time is similar, except that the directions of the circumferential components of normal force and stress are opposite. In dynamic During the process, the blade adaptively deformed elastically, and the two sections of the wheel hub moved relative to each other. This adjustment movement can also be done quickly under the hindrance of small mechanical friction, because small disturbances and vibrations are always present enough to help overcome friction.

 The radial inflow pre-rotator is mainly installed in the center position of the semi-open centrifugal guide wheel, which is closely connected with the suction chamber of the next stage impeller. In fact, the combination of the centrifugal guide wheel and the pre-rotator makes the latter stage impeller need no suction chamber at all, and this design is embodied in the multi-stage pump embodiment of the present invention.

 Referring to FIG. 7, as an example of application as a pre-rotator, the figure shows a semi-open type high potential ratio impeller structure with a pre-rotator as a combination of an axial inflow pre-rotator and a high potential ratio impeller. An embodiment. Among them, 31 is a semi-open high potential ratio impeller, 32 is the lower rim of the pre-spinner, 33 is a sail blade of the pre-spinner, 34 is a rigid rib of the pre-spin, and 35 is an upper wheel of the pre-spin .

 As shown in the figure, the pre-rotator is installed in the suction chamber of the high potential ratio impeller 31. The suction chamber is a circular cylindrical shaped cavity outside the impeller's central sleeve and inside the blade root. The two sections of the pre-spindle 圏 32 and 35 sliding sleeves (dynamic fit) are on the impeller shaft sleeve, and the four rigid ribs are radially fastened to the entrance surface of the root of four of the eight impeller blades. The assembly is complete. In the figure, the number of blades of the pre-rotator 4 is a sub-multiple of the number of blades of the impeller 8.

 High potential ratio impellers equipped with pre-spinners can be used in single-stage pumps or multi-stage pumps and can be assembled into centrifugal pumps as a whole. The pre-spinner is a supporting component of the design of the 90-degree inlet angle of the impeller flow channel of the present invention, and its adaptive pre-spin mechanism can play an important role in improving the inlet hydraulic characteristics and anti-cavitation characteristics of the pump, especially when the pump runs away from the design conditions. At this time, its adaptive mechanism is of special significance for improving the operating efficiency of the pump and extending the service life of the pump.

Referring to FIG. 8, as an example of an average speed bifurcation technology that suppresses relative vortex and comprehensively improves the operating characteristics of an impeller, the schematic structure of a high potential ratio impeller with a uniform speed bifurcation is shown in the figure. Among them, 36 is an L-shaped blade, 37 is a constant speed combing blade, 38 is a near-pressure side bifurcation, 39 is a near-suction surface bifurcation, 40 is a branch exit, and 41 is a branch entrance , 42 is a blade groove acceleration area near the branch exit, and 43 is a unilateral constraint speed finishing area outside the blade groove outlet.

 As shown in the figure, the number of L-shaped blades is arranged on the circular or zigzag wheel of the high potential ratio impeller in a rotationally symmetrical manner. In the middle of the front of the channel of the flute between the L-shaped blades, there are 2 blades. Combining leaves at an even speed produces 3 average speed forks. 38 is a near-pressure surface bifurcation, 39 is a near-suction surface bifurcation, and there is an intermediate branch between the two. The front and middle parts of the uniform-speed combing leaves also show a radial trend, the tail turns smoothly, and the downstream direction points to the accelerating section of the blade groove. Comb leaves form dense constraints on the flow under the premise of relatively low velocity. The integral of the normal force along the way will include a shear moment, which acts in the opposite direction to the direction of the relative vortex. One of the important factors. The more important mechanism is that, because the bifurcation entrance approaches 41 as shown in the figure and does not reach the leaf trough entrance, the bifurcation exit does not reach the leaf groove as it approaches 40 as shown in the figure, and its exit cross-sectional area is based on an experience obtained through an optimized test Coefficient distribution, its near-pressure side bifurcation distribution is more and near-suction surface bifurcation allocation is less, so the former along the way pressure is lower all the way and the latter along the way pressure is higher all the way, the difference of this pressure field gradient distribution, in the relative speed On the lower premise, it is the resistance to resist relative vortices, which is the main mechanism for curbing vortices. In addition, the pressure difference is transmitted to the inlet area with a small coefficient, and the pressure gradient formed is a containment factor for the vortex in the inlet area without combing leaves. The difference in the exit speed caused by the pressure difference can form an ejection force to accelerate the bifurcation exit near the pressure surface as shown in Figure 42 and accelerate it, and the speed gradient formed outside the outlet as shown in Figure 43 reduces the absolute speed and inside of the outer side. The relative velocity, which is the ideal low-loss velocity distribution. These two factors coupled with the favorable factors such as the longer bifurcation flow near the suction surface, the relative vortex can be contained under the optimized ratio of the exit area of the bifurcation.

Without relative vorticity, the relative velocity in the lobes and their forks will be reduced by half an order of magnitude, which is particularly important. Under the normal size impeller size and speed, the negative superposition of the relative vortex on the pressure surface and the positive superposition on the suction surface may cause the liquid flow velocity to exceed the critical value and enter a turbulent state. After reducing by half an order of magnitude, the flow velocity can be PT / CN2004 / 001413

The design is within the hydraulic specifications. At this time, the process loss or local resistance loss is so small that it can be ignored. As mentioned above, the design premise of relatively low speed can obviously be completely satisfied.

 The high-potential bifurcation with average speed bifurcation has the obvious advantages of high hydraulic efficiency and good cavitation resistance. The combination of this technology with design features such as semi-open or closed, single-stage or multi-stage, with or without pre-spinner, with or without serrated wheel, with or without internal friction reduction device, has no contraindications, so it can be widely used application. Technically, the structure shown in the figure is easy to manufacture. The simplest process is molding, including die-casting, powder metallurgy, injection molding, compression molding and other process routes, and only the cheapest two molds are used. Its production cost is very low, and its dynamic balance characteristics are good.

 Referring to FIG. 9, an embodiment of a combination of a uniform speed and high potential ratio impeller (hereinafter referred to as a high potential ratio impeller with a uniform speed bifurcation, the same hereinafter) and a prerotator is shown. Among them, 44 is an L-shaped blade of an impeller, 45 is a uniform speed fork, 46 is a lower wheel of a pre-spinner, 47 is an upper wheel of a pre-spinner, 48 is a pre-spinner sail blade, and 49 is a pre-spin Of the rigid ribs, 50 is the impeller shaft sleeve, and 51 is the impeller shaft.

The impeller with average speed and high potential ratio is assembled and positioned on the rotating shaft 51 through the sleeve 50. The impeller has L-shaped blades such as 44 to form the same number of blade groove channels. Each blade groove flow path is provided with a uniform speed combing leaf, forming a uniform speed bifurcation such as 45 and so on. The periphery of the impeller shaft sleeve is a circular cylindrical suction chamber cavity, and an axial inflow pre-rotator is installed in between. The lower rim 46 of the pre-rotator is installed in the lower part of the suction chamber, and the upper wheel 圏 47 is installed in the upper part. They are all made of self-lubricating materials such as polytetrafluoroethylene, so they can be independently rotated on the impeller shaft sleeve. The outer surfaces of the two rims are rotating curved surfaces that meet each other, and the direction angle of the generatrix is continuously rotated through 90 degrees, so that the fluid flow completes the conversion from the radial motion component to the axial motion component during the loading rotation. As shown in FIG. 48 and the like, the elastic sail blades of the prerotator are suspended on rigid ribs such as 49 in the figure, and the ribs are fixed at the root entrance of the impeller blade. During operation, the ribs rotate along with the impeller, which drives the sail blades and the two rims to rotate synchronously. The sail blades perform work on the liquid flow to pre-spin. Distributed deformation of the blade due to distributed load, The magnitude and direction of its equilibrium stress make the blade a propeller-shaped surface, and thus determine the lag angle of the rim. The propeller-shaped surface and the wheel lag angle are adaptively adjusted when the fluid flow changes dynamically. This mechanism can reduce the loss, including the contribution of the adaptive change of the blade angle of attack.

 When the L-shaped vane high potential ratio impeller is applied to an axial inflow centrifugal pump, the pre-spinner is an important configuration, so this combination will be a commonly used design. Among them, the zigzag disc half-open average speed high potential ratio impeller can be used to assemble a high potential ratio single-stage pump without an internal friction reduction device, which is the simplest of the conventional high potential ratio single stage pump .

 Referring to Figure 10, a more complex and more efficient single-stage high-potential ratio single-stage pump is shown. Among them, 52 is a volute, 53 is a trapezoidal groove guide ring, 54 is an average speed high potential ratio impeller, 55 is a pre-spinner, 56 is a mechanical shaft seal, 57 is a cantilever shaft, 58 is a rear cover, and 59 is a front cover. .

 The pre-spinning average speed high potential ratio secondary worm cantilever pump is composed of a closed-type high-speed potential ratio impeller 54, a pre-spinner 55 installed in the impeller suction chamber, a mechanical shaft seal 56, and a curved surface with a secondary worm 52 The front cover 58 and the rear cover 59 and the cantilever shaft are composed. Among them, the impeller 54 is a semi-open type or a closed structure with a riveted front cover, and 8 L-shaped blades separate 8 blade groove flow channels. Each blade slot has 2 uniform-speed combing leaves, forming 3 uniform-speed bifurcations. When the impeller works, it outputs a high potential-to-specific liquid flow. The pre-spinner 55 is installed in the impeller suction chamber, and has 4 elastic sail blades, which are fixed on 4 rigid ribs and 2 rims, and the rim slides on the impeller sleeve.

 This impeller has a high output potential ratio, anti-vortex, anti-cavitation, and adaptability to changing conditions. Among them, the pre-spinner pre-spins the liquid flow, which can avoid the impact of turbulence at the entrance; the high potential ratio design of the L-shaped blade flow channel and its outlet can reduce the output speed; the uniform speed bifurcation blocks the relative vortex and can prevent backflow, The trailing edge vortex and other harmful flow patterns make the flow field low-speed laminarization, and the resulting speed gradient can also reduce the wheel friction speed and the absolute speed in the vicinity of the exit.

The liquid flow at the outlet of the high potential ratio impeller has a uniform distribution effect that simulates continuous openings. The radial component of the mouth flow velocity is equal to the flow rate divided by the area around the wheel, which has a lower spicy value, which is conducive to the use of worm diversion. However, when the width of the impeller outlet is small, such as using straight worm guides or wide rectangular cross-section guide rings, local excitation will occur. In this example, a secondary volute or a small-entry guide ring combined with a secondary volute is used to design the diversion.

Among them, the secondary worm trajectory 52 adopts an optimized cross-section design, and is formed by smoothly matching two cross-section sections. The initial section is a long semi-elliptical section with a fixed long axis 2L, and the short semi-axis is b = Q0 (2α-1 + (1-α) θ / π) / ( ² LY) ·. (9) After becoming a semicircle, it is a large bow with a fixed chord length of 2L. The radius of the bow r is constrained and exceeds the equation r ^ TT-Sin— ^ L / r)) + L ₂ (r ² -L ² ) ^0,5 = QG (2α-1 + (1-α) θ / π) / (2πν) ... (10)

 Here, Q, θ, α, L, and V are the design volume flow, the center angle of the worm section, the worm preference coefficient, the half of the worm inlet cylinder width, and the average worm outlet velocity. The minimum friction surface and minimum gradient design method and its equations (9) and (10) are of great significance for the design of low-loss worms, where (10) can be analyzed by power series of order 2 or higher, or solved by numerical solutions . The worm optimization coefficient α is a true decimal value between 0.5 and 1. The specific value can be derived through theoretical analysis, or it can be obtained through a small number of experiments by using the preferred method.

 (9), (10) In the formula, the width of the worm entrance is the width of the impeller exit. When the width is small, local excitation may occur, and an isosceles trapezoidal groove guide ring (53) should be inserted for transition. The width of the bottom edge of the entrance and exit of the trapezoidal section is equal to the width of the entrance of the worm and the exit of the impeller. The ratio of the two is

Among them, ω, R, and K are respectively half of the width of the impeller exit cylinder, the impeller angular velocity, impeller radius, and impeller feedback reduction ratio. The height of the trapezoid is generally 3 to 5 times the width of the impeller exit cylinder, which is 61 ^ ~ 101 ^ More appropriate. Obviously, the guide ring is used as a transition flow channel, and its parameters are restricted by the parameters of the front and rear flow channels. After inserting the guide ring, the impeller output liquid flow enters in a nearly tangential direction, and its streamline will be continuously turned into a logarithmic spiral line close to the arc due to the radial pressure gradient distribution in the guide ring. Its radial coordinates Increment means that the cross-sectional area of the fluid flow is enlarged, and it also means that the kinetic energy must be consumed to enter the high-pressure area to overcome the pressure difference. This is the trapezoidal groove guide ring pressurization process. The guide ring set to eliminate local excitation will first It can automatically undertake the task of diversion and boost pressure, and then it can complete the task without local excitation and communication with the manifold. Inserting the guide ring can reduce the load of the worm, its optimization coefficient α will increase, its speed gradient will decrease, and the work efficiency will be improved accordingly. When α is close to 1, the volute will mainly function as a manifold. Therefore, the diversion load ratio of the guide ring and the worm optimization coefficient α are interrelated.

 The front and rear casing covers (59) and (58) of the pump are typed on the symmetric surface of the volute, and can also be typed on the cylindrical surface of the worm inlet at the front end. The latter parting process helps to reduce the volume. The shell cover is manufactured by precision casting or molding process, and its outer side has a radial rib structure to increase strength and save materials.

 As can be seen from the figure, the volute in this example occupies a large space, which is the result of long distance running and confluence. The disadvantages of this pump structure are obvious. The large size is not only inconvenient to use, but also increases manufacturing costs. This structure is not the best embodiment of the present invention because it follows the frame of the traditional single-stage pump worm. However, the high-efficiency feature of the single-stage pump design using the worm is very prominent.

 Referring to FIG. 11, a flow channel structure of a two-phase flow cooling shaft seal driven by friction reduction in a shaft-end cavity is shown. Among them, 60 is a rotating shaft, 61 is a shaft seal cavity structure, 62 is an annular static static seal, 63 is an annular static friction plate, 64 is a two-phase inflow pipe, 65 is an annular dynamic friction plate, 66 is an annular rotating static seal and its Compression spring, 67 is a circlip or bayonet, 68 is a ring-shaped cover plate, and 69 is a ring-shaped ring-shaped opening.

Schematic structure according to mechanical dynamic seal. The rotating shaft 60 protrudes from the shaft seal end casing, and the annular shaft seal static friction plate 63 is fixed on the shaft seal cavity casing through a static static seal 62 and forms a cylindrical surface seal between the casing and the moving friction plate. Ring-shaped moving friction plate 65 On the other hand, a cylindrical surface and a flat static seal are performed by a ring-shaped rotary static seal and its compression spring 66, and the pressure and torque transmitted by the circlip or the bayonet 67 are transmitted to rotate synchronously with the shaft. A dynamic seal is formed between the moving friction plate and the static friction plate, and a certain amount of heat is generated. If the heat power is not transmitted, the dynamic seal and its connecting member will be burned due to continuous heating.

 In the figure, the internal antifriction drives the two-phase inflow pipe 64 into the shaft seal cavity facing the static friction plate, and the annular cover plate 68 separates the shaft seal cavity from the adjacent end cavity, leaving a ring-shaped annular opening 69 communicating with the end cavity. . The two-phase flow will impact the liquid flow of the static friction plate and disturb the bubble floating, which can enhance the surface liquid flow velocity. The two-phase flow is rotated and separated by gas and liquid under the driving of three rotating parts 65, 66, 67. The radial coordinate of the liquid level is the same as the opening 69, and the cavity is almost filled with the rotating liquid ring. There is relative movement between the liquid ring and the friction plate, and its thermal convection circulation mechanism is beneficial to reduce the temperature around the heat sink. After the liquid absorbs heat and heats up, the two-phase flow is recombined with the separated gas at the outlet, and is sprayed to the end face of the impeller into the end cavity at a lower axial velocity.

 Internal friction reduction is an important design of the present invention to overcome the friction loss of the disk and improve the internal mechanical efficiency. Shaft seal cooling involves the safe operation of the pump and is a necessary technical design. Compared with the separate cooling cycle of the organization, the method of the present invention integrates the end-cavity two-phase flow cycle and the shaft seal cold cycle into a single tube design, saves the circulation flow of the pressure liquid, and can thus improve the volumetric efficiency of the pump. In addition, the viscosity coefficient of the two-phase flow after heating is reduced, which is beneficial to increase the friction reduction effect.

 Referring to FIG. 12, there is shown a gas barrier gap structure that prevents gas from escaping when the front end cavity is antifriction. Among them, 70 is the impeller cover, 71 is the front cavity, 72 is the inlet dynamic fit clearance and centrifugal flow channel, 73 is a small moving ring fixed to the near-axis portion of the impeller front cover, and 74 is the dynamic fit clearance and leakage flow channel. , 75 is a two-phase inflow pipe, 76 is a stationary cavity wall of a front-end cavity or an inlet pipe of a pump, and 77 is a suction chamber or a through flow channel thereof.

In the figure, an annular groove is arranged at the near-axis portion of the cavity wall of the front end cavity, and a small moving ring 74 rotating with the impeller is included in the groove, separating the annular groove into a V-shaped annular groove with a distal far end, and an annular gap 72 on one side communicates with the front end The cavity 71 and the annular gap 74 on the other side communicate with the suction chamber. N2004 / 001413

The two-phase inflow tube communicates with the gap 72, and the two-phase flow is centrifuged under the driving of a moving ring. Part of the flow of liquid turns from the gap at the bottom of the V-shaped groove to 180 degrees and slowly flows into the gap 74 and the suction chamber 77. Its flow resistance has a blocking effect. Can prevent gas from escaping, thereby forming a V-shaped air gap. The paraxial space where the gas floats up to the gap 72 flows into the front-end cavity 71. The liquid flow in the two-phase flow should be greater than the flow flowing through 74, and the remaining portion also enters the end cavity 71 through the gap 72, and then flows into the deflector from the impeller edge gap. The above-mentioned selective separation mechanism only requires a small component size, because the design goal is only to make the leakage side fluid be pure liquid in a small neighborhood of the branch.

 The tiny circulating flow inflation design of the impeller front cavity relies on the normal operation of the choke gap. In the design of the choke gap, the flow from the annular groove 74 to the suction chamber should have a large resistance coefficient, and the liquid flow of the two-phase flow should be greater than the leakage flow of the annular groove, which are two important prerequisites. If this premise cannot be met, the former will result in an increase in the circulating flow of the two-phase flow, and the latter will cause avalanche-like pressure relief in the aerated end cavity to be filled with liquid, and the state of aerated friction reduction will no longer exist.

 If some volumetric efficiency indicators are sacrificed, such a complex air gap may not be used, and only the flow rate of the two-phase flow needs to be increased, allowing the two-phase flow to be directly shunted back to the suction chamber, and internal friction reduction can be achieved to obtain efficiency. Of improvement. This is because: First, when the two flow channels that are diverted maintain a two-phase flow with the same gas-liquid ratio over a certain bifurcation length, the flow will be stable, and there is no possibility of avalanche gas escape Second, the change rate of volumetric efficiency to the loss of two-phase flow is a small constant, and the change rate of mechanical efficiency gain to internal friction reduction is a step with a relative amplitude of 90% or higher, and its amplitude is High power functions have high sensitivity, so sacrificing the former can obtain efficiency gains.

Referring to FIG. 13, there is shown an embodiment of an inflatable driving device of a semi-open impeller cantilever pump and its connection. Among them, 79 is the pressure liquid flow regulating valve, 80 is the intake flow regulating valve, 81 is the ejector, 82 is the mechanical shaft seal of the pump, 83 is the shaft cover cavity annular cover outlet, and 84 is the rear of the semi-open centrifugal pump. End cavity. In the figure, the pressure liquid enters the ejector 81 after adjusting the flow rate through the adjustment valve 79, and the gas adjusted by the adjustment valve 80 is ejected to generate a two-phase flow, which is connected to the shaft seal cavity through the pipeline, and after cooling the mechanical shaft seal 82, the ring cover The central outlet of the plate flows into the rear cavity to inflate it to reduce friction. The pressure liquid driving the ejector can generally be diverted from the outlet pipe of the centrifugal pump, and the pressure can be normally operated when the pressure is higher than the static pressure of the impeller by more than 0.05 MPa. The operating parameters of the centrifugal pump can usually meet this condition. The regulating valve 79 is used to regulate the flow of the pressure liquid, and is set to a small flow that can be normally inflated. The regulating valve 80 is used to regulate the gas flow rate. An incorrect opening degree may cause the output pressure to be too low to inflate the end cavity, or the inflation diameter ratio cannot reach the maximum value, so it must be capable of fine adjustment. The gas used should be harmless to the pumped liquid. Since the ejector has a wide range of pressure requirements on the gas source, air can be used as the inflation medium in most cases. At this time, the inlet of the valve should be connected to the atmosphere. The capacity, ejection pressure ratio, and output pressure of the ejector 81 should match the maximum inflation flow and the maximum end cavity pressure required for friction reduction in the end cavity, otherwise the expected effect cannot be achieved. The gas in the inflatable end cavity is generally not consumed, so the flow required by the entire device is small.

 Referring to Fig. 14, the figure shows an inflatable friction reducing device of a closed impeller cantilever pump and a connection scheme thereof. Among them, 85 is the mechanical seal at the end of the shaft, 86 is the rear cavity, 87, 88 are the rear cavity, the front end cavity flow distribution tube, 89 is the front cavity, 90 is the ejector, 91 is the ejector gas flow adjustment The valve, 92 is a pressure liquid flow regulating valve, and 93 is a front-end cavity V-shaped air gap.

The internal friction reducing driving device of the closed impeller centrifugal pump is composed of a pressure liquid regulating valve 92, a jet 90, an ejection gas regulating valve 91, flow distribution pipes 87 and 88, and a front-end cavity air-blocking gap 93. When the outlet pressure of the pump is higher than the static pressure of the impeller by more than 0.05 MPa, the pressure liquid is diverted from the outlet. 92 is used to adjust the pressure liquid flow, set to a smaller flow that can be normally inflated. 91 is used to adjust the gas flow, and its inlet is open to the atmosphere when air is used. The flow rate, maximum pressure and ejection pressure ratio of the ejector are matched with the required flow, maximum pressure and inlet pressure. Unlike the semi-open impeller pump in Figure 13 01413

Similarly, both end cavities of the closed impeller need to be inflated, and the front end cavity has a parallel return gap through the suction chamber, which needs to be set as a V-shaped air gap as shown in FIG. 12. Because the front and rear cavities are inflated at the same time, the flow of the two-phase flow is larger. The function of the delivery flow distribution pipes 87 and 88 is to set the difference in the pipeline resistance coefficient to control the flow distribution, provided that the parallel target end cavities have the same pressure in the inflated state. Fortunately, this is exactly the case, so during steady state operation, the flow will be controlled by the distribution pipe. During the initialization dynamics, the front-end cavity will be inflated first, and then the rear-end cavity.

 In the figure, the choke gap 93 is a V-shaped ring groove, which is used to separate and split the two-phase gas and liquid. The small two-phase ring that rotates with the impeller rotates the incoming two-phase flow. The centrifugal force causes the gas to float to The paraxial space is isolated by the small moving ring, and the liquid bypasses the moving ring from the bottom of the ring groove far shaft and enters the leakage gap, thereby blocking the gas leakage channel.

 In the alternative to the V-shaped groove air-blocking gap, in addition to the above-mentioned simple method of increasing the two-phase flow rate, a soft stop ring made of rubber, polytetrafluoroethylene, nylon and other materials can be used at the same position instead of V. Slots block flow. Under the premise of liquid lubrication, they can form a small and stable gap with the rotating shaft, the resistance coefficients to liquid and gas are sufficiently large, and the flow loss of the two-phase flow will be sufficiently small. The specific method is to open the outlet of the two-phase inflow pipe facing the contact between the soft stop and the shaft. Under the drive of the pressure difference between the end chamber and the suction chamber, the soft stop will be well lubricated.

 Referring to FIG. 15, the figure shows a schematic structure of a centrifugal guide wheel, which is a component with an assembled housing. Among them, 101 is a circular cylindrical shell (with a nose-shaped fastening bolt through hole), 102 is an axial projection of the flow channel of the transfer section, 103 is a guide wheel base plate, 104 is a guide wheel bushing, 105 is a bolt hole, and 106 is Guide vanes with a decreasing radius of curvature, 107 is the flow path of the transfer section from the impeller to the guide wheel, 108 is the cut-off tongue of the transfer section flow path, 109 is the deceleration boost flow path, and 110 is the cylindrical exit of the flow path.

The centrifugal guide wheel is composed of a circular cylindrical housing 101, a base plate 103, a shaft sleeve 104, and a guide vane such as 106 with a gradually decreasing curvature radius. The casing is provided with a transfer passage cavity such as 107, and a deceleration and pressurization passage such as 109 is provided between the guide vanes. The number of guide vanes of the centrifugal guide wheel is less than that of the impeller The number of blades is because the absolute flow velocity in the guide wheel is much higher than the relative flow velocity in the impeller, and a larger equivalent diameter is required to reduce the drag coefficient. But the number of guide vanes should not be too small, too few will prolong the confluence process of liquid flow and increase the friction area. In practice, the optimal number of guide vanes at the design size can be determined through theoretical planning or optimization experiments (using the optimization method). In the flow channel of the transfer section, it should be ensured that the liquid flow has no sudden change in speed and direction. In this way, the axial displacement will be completed at a low speed, the peripheral speed will be basically maintained, and local resistance losses such as commutation losses will be greatly reduced.

The design of the cross-sectional area expansion rate of the supercharged runner section of the centrifugal guide wheel and its change is more complicated. As a mechanical parameter, it should increase or decrease with the size of the viscosity coefficient of the pumped medium, because it is one of the key factors determining the kinetic energy loss rate ξ ₂ of the resistance type diversion along the way. As a geometric parameter, it is convenient to define the expansion rate as a function of the center angle of the diversion circle. The distribution of the function value is determined by the distribution of the radius of curvature of the guide vane at the boundary of the flow channel. Considering the symmetry of rotation, this is the distribution of the radius of curvature of the two relevant points before and after the same guide vane, and the width of the flow channel is intuitively reflected along the way. The accumulation of the differences in the radius of curvature of the relevant point groups, where the distance of the relevant points is related to the number of flow channels and changes with the change of the center angle of the diversion circle. Since the overall reduction ratio is equal to the inverse ratio of the cross-sectional area of the inlet and outlet, according to the distribution of the expansion rate and the inverse ratio, the differential point method is used to directly calculate the relevant point group and its distance to the centerline of the relevant flow channel, and the vane coordinates can be calculated point by point. , So as to accurately design the required shape of the guide vane. A constant expansion rate can be used in the design, but the optimization scheme is based on a variable expansion rate.

The outlets of the guide runners converge on the outer cylindrical surface of the central ring cavity of the guide wheel. The circular cylindrical area between the cylindrical surface and the shaft sleeve is the exit convergence area of the guide wheel. As shown in the figure, the outer surface of the shaft sleeve of the guide wheel is a rotating curved surface that rotates the fluid by 90 degrees in the axial direction. A radial inflow pre-spinner as shown in Figure 6 can also be installed in the circular cylindrical exit confluence area of the guide wheel. The pre-spinner has a speed field arranging function, which can change the direction of the axial speed at the same time as the liquid flow rotates And the size, make it turn 90 degrees as a whole to output from the axial direction, the installation of the pre-spinner can improve the efficiency and improve the operating characteristics of changing conditions. The structure of the centrifugal guide wheel is particularly suitable for manufacturing by a two-clamp molding process, and the cost of mass production is very low.

 Referring to FIG. 16, the schematic structure of the flow path of the transfer section of the combination of the high potential ratio impeller cavity and the centripetal guide wheel is shown. Among them, 111 is a region on the outer shell of the guide wheel constituting the transfer section flow channel cavity and its cavity wall support, 112 is a section of the transfer section flow channel guide wheel portion, 113 is a section of the transfer section flow channel impeller cavity portion, and 114 is Impeller cavity cover, 115 is the rear bottom face of the guide wheel, 116 is the depth position of the runner at the beginning of the new cycle, 117 is the position where the runner is axially located at the bottom plane, 118 is the two-section cross-section boundary line and the cutoff tongue appears Position, 119 is the minimum marginal curve of the impeller cavity section, and 120 is the maximum marginal curve of the impeller cavity section.

 The enlarged part in the figure indicates the schematic contour of the cross section of the flow channel in the transfer section. For the sake of simplicity, the change in radial coordinates is not considered. The flow channel of the transfer section of the centripetal guide wheel starts from the cut-off tongue of the previous diversion runner inlet section (its axial plane is projected as line segment 118), and ends at the cut-off tongue of the connected diversion runner inlet section. The flow path spans the impeller cavity and the guide wheel, and is divided into an impeller cavity part and a guide wheel part. The former is a confluence space surrounded by the impeller outlet cylindrical surface and the impeller cavity cover. The front cover is determined along the curved surface outside the cavity wall surface. The latter is the inlet section of the diversion channel, which runs through the impeller cavity. The two parts are assembled and joined together, and the shape and area of the composite cross section change periodically with the change of the center angle of the guide wheel. The rules are: a. From the start point to the end point, as the diversion center angle increases, the area of the composite section increases linearly from the minimum to the maximum. Its proportionality factor is equal to the design value of the volume of the liquid flow discharged by the impeller through a unit angle divided by the design value of the absolute speed of the liquid flow outlet, or it is multiplied by an expansion coefficient greater than 1 and less than the minimum expansion ratio of the booster flow path of the guide wheel, so The transition section runner also has a deceleration and boost function.

Among them, the cross-sectional area of the two parts is changed in two sections. From the starting point to the positive position of the pressure-increasing section of the previous diversion runner, that is, the rear end of the section is moved to the point (angle) at the bottom plane of the runner as shown at 117 in the figure, and the cross-sectional area of the impeller cavity linearly increases from the minimum value to The maximum value, the cross-sectional area of the guide wheel part remains at 0. From this point to the end, the impeller cavity part 001413

The cross-sectional area decreases linearly from the maximum to the minimum, and the cross-sectional area of the guide wheel linearly increases from 0 to the maximum.

 b. The minimum value of the composite cross-sectional area is the same as the minimum cross-sectional area of the impeller cavity, which is equal to the area of the curved triangle formed by the straight line segment 118, the curve 119, and the cylinder generatrix of the impeller. Derived from the limit value of the axial acceleration. Curve 119 is composed of two elliptical arcs and the middle arc. The analytical parameters are determined by the cross-sectional area and the coordinates of the endpoints. The maximum value of the composite cross-sectional area is equal to its minimum value plus the maximum cross-sectional area of the guide wheel part. The latter is equal to the product of the composite cross-sectional area increase ratio coefficient and the center angle that is bypassed from the start point to the end point. The maximum cross-sectional area of the impeller cavity is equal to the area of the curved triangle formed by the straight line 118, the curve 120, and the generatrix of the impeller's cylindrical surface. Curve 120 is the maximum margin of the cross section of the flow channel during the cut-off period of the flow channel inlet section. It is formed by two elliptical solitary kisses. Its analytical parameters are the increase in the combined cross-sectional area of the end point and the flow channel during the cut-off period. determine.

 c. From the start point to the end point, as the diversion center angle increases, the shape of the cross section of the two sections changes in two sections. During the increase of the cross-sectional area of the impeller cavity, the cross-sectional shape is a curved triangle, and its curved edges start from the position of curve 119 and change to curve 120 through a series of intermediate processes similar to a curve. During the increase of the cross-sectional area of the guide wheel, the cross-sectional shape starts from the initial straight line section 118 and undergoes various changes. First, the long semi-ellipse with 118 as its long axis gradually increases its short semi-axis; after it becomes a semicircle, Instead, the semi-circle margins are continuously moved forward to form a rectangular cross section followed by a front semi-circle. When the tongue appears, the semi-circle moves to the 116 position, and the transfer section flow path is separated from the impeller cavity to become a pressurized flow path.

d. From the appearance of the tongue, the flow channels defined by 116 and 118 continue to move forward and deform. After a small amount of forward movement to form the minimum physical width of the tongue, the rearward side marginal line changes from the tongue's forward-to-margin straight line segment to a curved long semiellipse. The short semi-axis is continuously extended to become a semicircle and then changed to Pan until it reaches the bottom surface 117 of the runner. In this process, its forward marginal semicircle moves forward continuously until it is tangent to the bottom plane 115 of the guide wheel, and changes to The continuous compression semicircle is a semi-ellipse, and finally becomes a straight line penetrating the bottom plane. During the section of the booster runner moving forward, its area changes according to the deceleration booster requirement.

 During the change of the cross section of the above-mentioned guide wheel part, the radial coordinate of its centerline may change. When the starting point of the guide vane is the starting point of the inlet section of the flow channel, the radial coordinate of the center line is continuously reduced. When the starting point of the guide vane is the end point of the inlet section of the flow channel, the radial coordinate of the center line is unchanged.

 The key to the design of the flow channel in the transfer section is to strictly control the changes in the cross section of the flow channel, including the changes in the shape and area of the two sections. Starting from the minimum cross-sectional area, through the process of variable cross-section confluence, combined variable cross-section confluence, and split cross-section of the impeller cavity, a method of constructing the flow channel cross-section margin and linearly expanding the cross-sectional area by using a long semi-ellipse to shorten the semi-axis, semi-circular translation, etc. Generate the required velocity field distribution and control the marginal friction loss index. The changes in the tangential, radial, and axial components of the flow velocity are derived from the impulse integral of the wall normal force and its dynamic impact on the pressure distribution. The influence of the velocity field distribution and its spatial change rate on the flow pattern of the flow channel is highly sensitive, and it has a significant impact on the efficiency. If it is not good, it will also produce water hammer chatter effect or cavitation cavitation effect. This is a complex problem of multifunctional curved surface design. The disclosed method is not yet optimal, but the goals of preventing destructive effects and controlling frictional areas and increasing the equivalent diameter have been reflected. Although the geometric planning techniques used make this design difficult to express and formulate processing technology using traditional methods, it is not difficult and cost-effective to implement using modern CAD, CAM technology and molding technology.

Referring to FIG. 17, a schematic structural diagram of a flow path of a closed impeller super friction reduction and guide wheel control transfer section is shown. Among them, 121 is the area on the outer shell of the guide wheel constituting the transfer section flow channel and its cavity wall support, 122 is the section of the transfer section flow channel section guide wheel, 123 is the assembly position of the front cover extended super friction reducing closed impeller, 124 is the impeller Cavity cover, 125 is the rear end face of the guide wheel, 126 is the forward side bottom of the transfer section when the diaphragm appears, 127 is the rear side bottom after the pressurized runner is aligned, 128 is the location where the tongue appears, 129 Is the cross section of the impeller cavity part of the transfer channel, 130 is the tail of the impeller blade, 131 is the impeller flow channel, 132 is the extended impeller cover, and 133 is the impeller cavity cover. 001413

The enlarged part of the figure marks the schematic marginal contour of the cross-section change of the runner in the control wheel transfer section. The guide vanes of such a centrifugal guide wheel have a starting point of variable curvature that is in contact with the circumferential cavity wall. The starting point is the positive position of the pressurized runner after the transfer, and the radial coordinate of the center of the runner in the transfer section is determined by this point. The cross section of the flow channel of the transfer section is divided into the impeller cavity part and the guide wheel part, and the two parts are assembled and joined together. The impeller cavity section of the cross section is wrapped in the impeller by the impeller cover, has a fixed area and shape, and it assumes the steering adjustment of the axial surface speed component. The guide wheel section of the cross section is a cross section of a space between two adjacent separating tongues which communicates with the impeller cavity. This section independently controls the process of confluence and tangential and axial movement. With the increase of the diversion center angle, the cross-section of the guide wheel changes periodically with the tongue as the starting point and end point. The change law in a cycle is:

 a. The cross-sectional area increases linearly from 0 to the maximum. Increasing the proportionality factor is equal to the design value of the volume of the liquid flow discharged by the impeller through a unit angle divided by the design value of the absolute speed of the liquid flow outlet, or multiplying by an expansion factor greater than 1 and less than the minimum expansion ratio of the booster flow channel of the guide wheel. The maximum cross-sectional area is equal to the increase of the proportionality factor multiplied by the center angle of the corresponding channel inlet section. When the scaling factor is increased to include the expansion coefficient factor, the flow passage of the transition section has a deceleration and pressure increasing function.

 b. The cross-section starts from the starting straight line segment 128 and undergoes various shape changes: First, a long semi-ellipse with 128 as its long axis, and its short semi-axis gradually increases; after it becomes a semicircle, it changes to a semicircle and the margins move forward continuously to form The front semicircle is followed by a rectangular cross section; when the tongue appears, the semicircle moves to the 126 position, and the flow passage of the transfer section is separated from the impeller cavity to become a pressurized flow passage.

c. After the diaphragm appears and isolates the impeller cavity, the flow channel defined by 126 and 128 becomes a pressurized flow channel, which continues to move forward and deform. After a small amount of forward movement leaves the minimum physical width of the tongue, the rear side margin changes from a straight line segment to a curved long semi-ellipse. The short semi-axis is continuously extended to become a semicircle and then translated to the final point. Enter the bottom surface of the runner until 127. In this process, its forward marginal semicircle continues to move forward until it is tangent to the bottom plane 125 of the guide wheel, and the semicircle is continuously compressed into a semiellipse, and finally becomes a straight line and 3

The bottom plane penetrates. The moving speed of the contour line or the center of mass should be greater than the appropriate moving speed during the confluence, such as more than 50%, so that the cross-sectional area and strength of the tongue can be continuously increased. During the forward-flow section of the booster flow channel, its area changes according to the deceleration boosting requirements.

 The complexity of the above scheme is significantly smaller than the design shown in Figure 16. Its velocity field distribution and marginal friction loss rate index are also better than the former, and the uniformity of the steering dynamic reaction force and wall normal force distribution of the axial component of the flow velocity is also Even better, the pressure pulsation is also small. This design is based on the premise that it is matched with the closed impeller with the front cover extended. When the end cavity is inflated to reduce friction, the friction reduction surface of the front cover side end cavity covers the entire friction surface of the impeller cavity portion of the flow channel of the transfer section. According to the fifth-order law calculation, it can obviously produce a large friction reduction benefit in the flow passage of the transfer section, so it is called super friction reduction. Ignoring the relative velocity friction on the inner side of the impeller cover, the friction surface of the flow channel of the transfer section will only have the guide wheel portion. Because the margin line of the partial section is not a closed curve, the optimal cross-sectional shape will no longer be circular. To optimize, you only need to make a plan to minimize the length of the non-closed margin line under the same cross-sectional area. The above scheme is a simple design with approximate optimization.

 The solution shown in Figure 17 also requires the use of 3D CAD to design a clearly expressed processing blueprint. The manufacturing cost of the molding process is also low, and the impeller cover is stamped.

Referring to FIG. 18, a schematic diagram of the structure of a symmetrical end cap of a centrifugal spur diverging angle outlet pipe is shown. Among them, 141 is the assembly stop of the end cover, 14 ² is the pressure-bearing cover plate, 143 is the deep entrance of the central worm channel, 144 is the shallow central worm channel, 1 ⁴ 5 is the bearing cavity supported by the runner envelope structure, 146 Is the axial projection of the central worm, 147 is the tongue between the deep and shallow parts of the central worm, is a radial straight line spanning the inner and outer marginal circle of the annular entrance, is the starting line of the worm, and 148 is the center In the shallow part of the worm, 149 is the sleeve and 150 is the outer circle of the annular entrance.

The symmetrical end cover module is composed of a pressure-bearing cover plate 142 with an assembly stop 141, a three-dimensional worm 144 on the cover plate and its annular entrance 143, and a straight line communicating with the worm interface 146. N2004 / 001413

It is composed of a segment pipe, a shaft sleeve 149 supported by a worm structure, and a bearing cavity 145. It is a multi-structure integrated part. The central worm channel of the end cap is a three-dimensional flow channel that synthesizes tangential, radial and axial movements. Its starting position is the tongue 147 on the circular plane of the circular inlet. The end position is increased in the radial and axial directions. Below the tongue of the coordinates. The entrance and exit of the volute and its interior have movement continuity in three directions. Its dynamic characteristics are that the time change rate of fluid acceleration is small, and the spatial change rates of flow field parameters and wall normal force are also small and constant. This is one of the mechanical characteristics of turbulent flow field stability. The geometric design due to mechanical considerations that are difficult to indicate in the figure also includes: Starting from the tongue, the cross-sectional area of the volute increases in proportion to the center angle, and the radial and axial coordinates of the center of the volute bottom gradually increase with the need to expand the cross-sectional area. Increased to form a three-dimensionally expanded snail slope. After one week of rotation, it entered the underside of the tongue, and then tangentially connected to the straight-line pipe. The shape of the worm section is as follows: the starting point is a straight segment of the tongue, and then the long axis becomes a semi-ellipse with a shortened axis on the entrance plane. After becoming a semicircle, it gradually sinks and smoothly increases the radius of curvature of the lower part. The involute curve with an appropriate curvature change rate develops until it enters the lower part of the tongue, and then maintains a cross-sectional area to deform into a circular cross-section to fit the pipe.

 Analyze and make full use of the geometric and mechanical characteristics of symmetrical end caps. Using the aforementioned modular combination method of the present invention, the following extensions of the target function or target use of the component technical design of the end cap will result in centrifugal pump design methods and applications. Many improvement possibilities in the way:

a. Utilizing the characteristics of the end cap annular interface and the internal compatibility and constraint of the three-dimensional motion of the three-dimensional volute, it is possible to construct or adaptively generate the boundary conditions of the shunt, confluence, rotation, and steering of the impeller and guide wheel in the multi-flow channel, so that Meet the requirements for the connection of the impeller inlet and the connection of the centrifugal guide wheel outlet, and it is universal for single-stage and multi-stage pumps, which has produced a flow field boundary module used as a single-stage and multi-stage pump Model, which has become the technical basis for supporting modular combinations and implementing conservative loop design; b. Exploiting the end cap annular interface and the directional compatibility inside the three-dimensional volute and the continuity of three-dimensional motion, and expanding it into the opposite direction of inflow and outflow, Reciprocal nature Compatible with the technical design, it can be used as the front and back flow symmetry boundary module, which becomes the boundary basis for the connection mode required to support the modular combination;

 C. Utilizing the continuity of movement in the three directions of the end cap annular interface and the three-dimensional volute, limiting and optimizing the spatial and temporal rate of change of the flow velocity to minimize it, which can be used as a flow with stability and low loss characteristics. Field boundary module to make the connection mode performance required by the modular combination better;

 d. Utilizing the rotational symmetry of the ring-shaped interface of the end cap and the assembly stop, and the relative positioning of the worm and the lead-out tube based on the relative angle of the tongue, so as to construct the functions of the front and rear cover independently variable angle outlet tubes to support the module The assembly structure design required for the assembly, and the connection relationship between the assembly and the actual fluid flow system;

 e. Utilizing the features of the integrated design of the pressure-resistant cover plate and bearing seat of the end cover, it is determined as a moldable single-part functional component in terms of technology and process design. After the expansion of the above technology is included, a new end cap module can be constructed that supports the connection mode of the assembly size and interface parameters can be standardized, in order to expand the value of its small size, simple design, low cost, and powerful functions.

 The technical expansion described above is just the right design. Matching the centripetal guide wheel with the symmetrical end cap, and then expanding it to the centripetal booster module and the variable angle outlet tube symmetrical end cover module, which can greatly reduce the volume of the centrifugal pump, and its variable angle outlet tube function is Users welcome that its symmetry, its technical design, and the increased versatility of its application methods will help simplify the design, manufacture, and use of centrifugal pumps. All of these features and feature expansions that take advantage of features help reduce both production costs and total cost of ownership for users. The advantages of this design are particularly evident when used in single-stage pumps. For example, in the cantilever pump shown in FIG. 10, FIG. 13, and FIG. 14, the size of the deflectors and the manifolds stacked on the periphery of the impeller is wasted, which makes people feel particularly unfortunate. By applying the design disclosed in this embodiment, this disadvantage can be achieved. avoid. From the description of the embodiments disclosed later, the application advantages of the symmetrical end cap module will be more clearly seen.

Figure 19 ~ Figure 27 are the axial combination of impeller and radial guide wheel structure according to the modular method An example of a centripetal booster module.

 Referring to FIG. 19 to FIG. 27, the common features and advantages of these impeller combination centrifugal booster modules are first comprehensively described, and then their individual characteristics and their effects are briefly listed in Table 10.

 These modules are energization modules containing different impeller technologies or processes. In accordance with the connection mode of "fluid flow in and out of the paraxial annular mouth with annular flow," one or more energization modules of the same specification are connected in series, and Combined with two symmetrical end cover modules of corresponding specifications, it can form different types of centripetal booster centrifugal pumps with module interchangeability.

 The centrifugal booster module is an axial combination of a centrifugal guide wheel, an impeller, and an impeller cavity cover. Some are also equipped with other functional accessories, with standardized interface parameters and assembly dimensions. Among them, the centrifugal guide wheel is integrally manufactured by molding, and the cavity side plane or rotating curved surface forms a clearance fit with the impeller. The outer edge of the cavity side has a forward marginal curved surface of the transfer section flow channel designed according to the impeller parameters. The casing has an assembly stop, and a middle partition plate integrally connected with the casing serves as a guide vane supporting substrate, and at the same time, it plays a role of isolating the impeller cavity and the guide cavity and withstanding the pressure difference therebetween; the impeller cavity cover plate is a molded weight-reducing structural part. The cavity-side rotating curved surface cooperates with the impeller to form a clearance, and the outer edge of the cavity side has a rearward marginal curved surface of the transfer section flow channel designed according to the impeller parameters. During assembly, the guide wheel, the impeller and the impeller cavity cover are sequentially installed, and the three are respectively positioned by the casing stop, the rotating shaft and the impeller cavity of the guide wheel. During operation, the liquid flow flows axially from the inlet of the module into the rotating impeller flow path, receives the normal force work of the blade to accelerate along the way and integrates the centrifugal force work to increase the specific energy, enters the guide wheel through the flow passage of the transfer section, and decelerates and boosts pressure in it. , Rotate 90 degrees out of the module from the paraxial annular opening with a loop.

Centripetal boost module is the product of modular combination method and conservative loop design. The centripetal guide wheel enables the centripetal booster module to have the spatial periodicity of the fluid flow process, the flow channel connection, and the flow parameter-the periodicity from the impeller inlet to the guide wheel outlet. This periodicity is the principle basis of the module division . This basis is the result of the purposeful design of the module for inversion requirements, not the use of accidental discovery. Conservative loop design aims to optimize flow field velocity The spatial and temporal rate of change in degrees comes from thinking about the macroscopic laws of local excitation and local loss of fluid machinery. The concept of complete constraints and respect for the moment of inertia of fluid momentum are the conceptual basis of conservative ring design. These thoughts and principles have been implemented in the aforementioned series of features of the present invention. They have good effects in optimizing the dynamic process of liquid flow transfer, stabilizing the flow field, and reducing losses, and even the dynamic reaction of the flow channel wall. Force loads can be reduced. Organizing and applying these designs to a modular approach with global value will be more effective.

 Based on the above ideas and technical routes, the present invention expands and plans the following common features and functional performance features for the centripetal booster module in the examples of Figs. 19 to 27: a. The centripetal booster module consists of various impellers and centripetals. The guide wheels are axially combined. The internal and external assembly dimensions and interface parameters are standardized, and their interchangeability covers the design, production, and use processes. This design can bring many benefits and conveniences for enterprises and users, and can greatly enrich and quickly spread the type of technology of centrifugal pumps.

 b> The flow path of the confluence and transfer section of the centrifugal guide wheel is connected in series rather than in parallel with the booster flow path, so there is no constraint conflict that takes into account the convergence and booster pressure. They all have a complete constraint wall surface, and the transfer and booster flow props It has a preset expansion rate, which can avoid under-constraint and local excitation phenomenon, and its boosting efficiency can be up to 98%. The peripheral speed of the impeller and the guide wheel is maintained at an overcurrent, so that the direction of the liquid flow is independent of the operating conditions. The speed matching characteristics of the impeller interface are good, and it has 100% adaptability to changing conditions. When the flow rate is reduced, the pump Instead, the efficiency increases.

c. Adopt the connection mode of "fluid flow in and out from the paraxial annular mouth with annular flow." The impeller diameter is the same, and the axial dimension of the impeller is smaller than the impeller. Therefore, the module has a compact structure and the smallest volume. The volume of the multi-stage pump is reduced by more than 50%, and the manufacturing cost is greatly reduced. The speed distribution of the impeller inlet and the guide wheel outlet are basically the same, and both have a large circumferential component. Such a design can not only improve the speed matching characteristics of the impeller suction chamber and adaptability to changing working conditions, but also reduce the diversion load and shorten the diversion process. Compared with the prior art, the former is a functional improvement that generates new features, and the latter is a structural improvement that improves performance. 04 001413

The overall effect of the user is to change the law of the efficiency curve, not just to improve the efficiency data by a point or a paragraph.

 d. As enumerated in the examples, the impellers in the module can combine different technical and technological characteristics, including high potential ratio, radial and axial pre-spinning, anti-vortex and equalization, internal friction reduction, and super-anti-friction. Innovative features and traditional back-bend designs also include semi-open and closed structural features. Combining modules with different impellers has different effects. The effects of improving the potential ratio and flow diversion efficiency, increasing the impeller stroke efficiency, improving the anti-cavitation characteristics, and improving the adaptability of variable operating conditions are directly derived from the innovative design of the impeller, which improves cost-effectiveness due to the modularity and mathematical planning described, And the many conveniences created by the interchangeability of changing combinations in the design, production, and use processes are the benefits of modular combinations, and their values vary by technology.

 The individual characteristics and effects of various centripetal booster modules are listed in Table 10. Table 10 Personality characteristics and effect description of centripetal boosting module

 Figure 19 In the Heart Report ^^ ^

Closed ^1- heart type, giving the ratio of extinction, in the miscellaneous 20 of the centripetal diagram ^ 鹓

 The following closed form gives a reduction of the ratio of b ^^ l to TO, Figure 21 The Hermetically Closed Form

 ^ ^ ^ 娜 ¾¾ · Τ¾Λ_ 管 等 柳 和 件 ， 充 ^ 行

~ 95%, ¾¾¾ rate is 5 ~ 9%. N2004 / 001413

Referring to FIG. 19, a centripetal booster module that axially combines a traditional semi-open impeller and a centrifugal guide wheel according to a modular method is shown. Among them, 151 is the impeller flow channel entrance, 152 is the impeller cavity cover, 153 is the assembly stop, 154 is the impeller, 155 is the impeller cavity portion of the transfer channel cross section, 156 is the guide wheel portion of the transfer channel cross section, 157 It is a centripetal guide wheel with a shell, 158 is a supercharged flow path of the guide wheel, 159 is a cylindrical surface of the guide wheel exit, and 160 is a rotating shaft that constrains the impeller in the module.

 In this example, the semi-open impeller centripetal booster module is formed by axially combining a centripetal guide wheel (157), a semi-open impeller (154), and an impeller cavity cover plate (152). Among them, the rotating curved surface on the guide wheel (157) and the impeller cavity cover (152), and the mating curved surface of the flow path of the outer edge transfer section of the both are specially designed according to the parameters of the semi-open impeller (154), including according to the shape and size of the impeller. Determine the generatrical coordinates of the rotating surface with the fit clearance, and determine the wall coordinates of the flow path of the transfer section according to the fluid parameters.

 This semi-open impeller centripetal booster module outputs a constant potential specific flow, which is suitable for assembling a centrifugal pump with an impeller speed of about 10 meters per second. The speed can be increased when the equivalent diameter of the flow channel is increased. Since the flow channel restriction degree of the transfer section is higher than that of the traditional deflector, the phenomenon of the return flow of the impeller outlet will be curbed. Its application advantages mainly lie in the technological benefits brought by the modularity and the functional and performance benefits brought by the combination.

Referring to FIG. 20, a centripetal booster module combining a classic closed impeller and a centrifugal guide wheel is shown. Among them, 161 is the inlet of the impeller flow path, and 162 is the impeller cavity cover. T N2004 / 001413

163 is the assembly stop, 164 is the traditional closed-type impeller, 165 is the impeller cavity part of the runner section of the transfer section, 166 is the guide wheel part of the runner section of the transfer section, 167 is the centripetal guide wheel with the shell, 168 is The guide wheel booster flow path, 169 is the cylindrical surface of the guide wheel flow path exit, and 170 is the constraining shaft of the impeller.

 In this example, the closed-type impeller centripetal booster module is formed by axially combining a centrifugal guide wheel 167, a closed impeller 164, and an impeller cavity cover plate 162. Among them, the rotating curved surface on the guide wheel 167 and the impeller cavity cover 162, and the mating curved surface of the flow path of the outer edge transfer section of the two are specially designed according to the parameters of the closed impeller 164, including determining the rotation according to the shape and size of the closed impeller and the fit clearance. The generatrical coordinates of the curved surface and the wall coordinates of the flow passage of the transfer section are determined according to the fluid parameters, and the space occupied by the impeller cover is excavated from the impeller cavity cover 162 so that the blade groove flow channel is in contact with the outlet flow channel. The assembly requirements, operating principles, adaptability, and application effects of this module are basically the same as the embodiment shown in Figure 19, except that the closed impeller has different performance.

 Referring to Fig. 21, a centripetal booster module combining a classic closed impeller with a centrifugal guide wheel and further combined with an internal friction reduction technology is shown. Among them, 171 is a V-shaped groove air-gap annular cover plate installed on the impeller cavity cover plate, 172 is a V-shaped groove small moving ring installed on the impeller cover plate and rotated one by one, and 173 is a V-shaped groove air gap clearance For the two-phase flow inlet, 174 is an inflatable drive two-phase inflow pipe that passes through the shell and the impeller cavity cover plate and enters the inlet 173, 175 is the impeller cavity cover plate, 176 is the front end cavity of the inflatable impeller, and 177 is arranged on the impeller to pass through The front and rear cavities in the blade's front cover fixing rivet communicate with pressure equalization holes, 178 is the aerated rear cavity, 179 is a guide wheel with a casing, and ΙδΟ is a closed impeller.

In this example, the anti-friction closed impeller centripetal booster module consists of a centrifugal guide wheel 179, a closed impeller 180 and an impeller cavity cover plate 175, and a V-shaped groove air gap annular cover plate 171, a V-shaped groove moving ring 172, Inflatable driving two-phase inflow pipe 174, front and rear cavity equalizing holes 177 in the front cover fixing rivet, and other internal friction reducing parts or structures are combined. Among them, the rotating curved surfaces on the guide wheel 179 and the impeller cavity cover 175, and the matching curved surfaces of the flow path of the outer edge of the transfer section are both It is specially designed in accordance with the parameters of the closed impeller 180, and includes digging out the space occupied by the impeller cover from the impeller cavity cover 175 so as to make the blade groove flow channel and the outlet flow channel match. The V-shaped groove air blocking gap formed by the ring groove on the impeller cavity cover plate 175, the ring cover plate 171, and the rotating movable ring 172 can maintain the internal friction reduction state. The pressure equalization hole 177 communicates the back-end cavity with the front-end cavity and reduces the friction of the isobaric filling gas without the need for additional takeover. The inlet tube 174 is connected to a two-phase flow driving pressure source such as a jet, and can drive friction reduction in the double-end cavity. The assembly requirements, operating principles, and adaptability of this module are basically the same as the embodiment shown in Figure 20. The combined internal friction reduction technology will reduce the disc friction loss of the closed impeller by 82%-95%, which will increase the pump efficiency. 5 ~ 9%.

 Referring to Fig. 22, a centripetal booster module of a modular combination of a half-open type average speed high potential ratio impeller and a centrifugal guide wheel is shown. Among them, 181 is the inlet of the impeller flow channel, 182 is the impeller cavity cover, 183 is the semi-open type average speed high potential ratio impeller, 184 is the exit turning acceleration section of the impeller flow channel, 185 is the tail of the L-shaped blade, and 186 is the transfer section flow The guide wheel part of the channel cross section, 187 is the guide wheel part of the flow path section of the transfer section, 188 is a centripetal guide wheel with a housing, 189 is a guide wheel pressurized flow path, and 190 is a cylindrical surface of the guide wheel flow path exit.

 In this example, the semi-open type average speed high potential ratio impeller centripetal booster module is formed by axially combining the centripetal guide wheel 188, the semi-open type high speed potential ratio impeller 183 and the impeller cavity cover plate 182. Among them, the rotating curved surface on the guide wheel 188 and the impeller cavity cover 182, and the outer curved surface of the flow transition mating surface of the two are specially designed to match the parameters of the half-open average speed high potential ratio impeller 183. During operation, the relative vortex is blocked by the uniform speed bifurcation, the liquid flow is laminarized at a low speed, and it accelerates in the acceleration section 184, and the outlet velocity decreases by the same amount.

The assembly requirements of this module are basically the same as the embodiment shown in FIG. 19. Due to the advantages of the half-open type average speed high potential ratio impeller, the impeller output potential ratio can reach 3 to 9, and the input speed ratio of the centripetal guide wheel will be greatly reduced. Therefore, the working linear speed of the impeller can be increased to more than 20 meters per second, and the speed can be selected higher when the equivalent diameter of the runner is larger. The advantages of uniform speed and high potential ratio impeller, as well as the process benefits and functional and performance benefits brought by the modular design, will be integrated in this example. 3

Reflected.

 Referring to Fig. 23, a centrifugal supercharging module of a modular combination closed-type high-speed average potential impeller and a centrifugal guide wheel is shown. Among them, 191 is a closed impeller with a uniform velocity and a high potential ratio, 192 is an impeller cavity cover, 193 is an impeller cover, 194 is an impeller cover fixing rivet, 195 is an impeller flow channel tail acceleration section, and 196 is an impeller cavity of a flow section of a transfer section. Part 197 is the guide wheel section of the runner section of the transfer section, 198 is the centripetal guide wheel with the housing, 199 is the guide wheel booster runner, and 200 is the cylindrical surface of the runner exit.

 In this example, the closed-type constant-velocity high-potential-ratio impeller centripetal booster module is formed by axially combining a centripetal guide wheel 198, a half-type uniform-velocity high-potential-ratio impeller 193, and an impeller cavity cover plate 192. Among them, the rotating curved surface on the guide wheel 188 and the impeller cavity cover 182, and the flow path mating surface of the outer edge transfer section of the two are specially designed to meet the parameters of the closed-type average velocity high-potential ratio impeller 193, including from the impeller cavity cover 192. Cut out the space occupied by the impeller cover, so that the flow path of the blade groove is in contact with the outlet flow path. During operation, the relative vortex is blocked by the uniform speed bifurcation, the liquid flow is laminarized at a low speed, and it accelerates in the acceleration section 195, and the outlet velocity decreases by the same amount.

 The assembly requirements of this module are basically the same as the embodiment shown in FIG. 20. Closed-type average speed The output potential-to-dynamic ratio of the high potential ratio impeller can reach about 3-9, and the impeller working speed is above 20 m / s, and the input speed ratio of the centrifugal guide wheel will be greatly reduced. The impeller range efficiency of this module is higher than that of the semi-open impeller module. Its average speed, high potential ratio advantage, and the process benefits brought by the modular design and the combined functions and performance benefits will be better reflected. This module is also the design basis for further combining other innovative technologies.

 Referring to Fig. 24, a centrifugal supercharging module of a modular combination closed-type average speed high-potential ratio impeller, an axial flow pre-spinner and a centrifugal guide wheel is shown. Among them, 201 is the axial inflow pre-rotator, 202 is the impeller cavity cover, 203 is the impeller cover, 204 is the closed-type average speed high potential ratio impeller, 205 is the tail acceleration section of the impeller flow channel, and 206 is the flow section of the transfer section. In the impeller cavity part, 207 is the section of the runner section of the transfer channel section, 208 is a centripetal guide wheel with a housing, 209 is a supercharging runner of the guide wheel, and 210 is an exit of the guide wheel.

In this example, the pre-spinning closed-type high-potential ratio impeller centripetal booster module is driven by a centripetal guide wheel. 13

208. Closed impeller 204 with uniform speed and high potential ratio, pre-rotator 201 and impeller cavity cover 202 are combined. Among them, the rotating curved surface on the guide wheel 188 and the impeller cavity cover 182, and the mating surface of the flow path of the outer edge transfer section of the two are specially designed to meet the parameters of the closed-type average high-potential ratio impeller 193, including from the impeller cavity cover 192. Cut out the space occupied by the impeller cover, so that the flow path of the blade groove is in contact with the outlet flow path. During operation, the relative vortex is blocked by the uniform speed bifurcation, the liquid flow is laminarized at a low speed, and it accelerates in the acceleration section, and the outlet velocity decreases by the same amount. The pre-spinner 201 generates constant velocity pre-spinning in the axial incoming flow, so that the flow field at the inlet of the impeller is adaptive to changes in working conditions and avoids turbulence and cavitation.

 The assembly requirements for this module are the same as the example shown in Figure 20. During operation, the impeller output potential ratio can reach 3 to 9, and the impeller speed is above 20 m / s. This module is more efficient than all the previous examples. The technological benefits of the modular design and the combined benefits of other modules will make the combined structure of the module of this example one of the hot selection components in the design of a centrifugal pump

 Referring to FIG. 25, a centrifugal supercharging module with a modular combination of a uniform average high potential ratio closed impeller, an internal friction reduction assembly, and a centrifugal guide wheel is shown. Among them, 211 is a V-shaped groove air gap ring cover plate installed on the impeller cavity cover plate, 212 is a V-shaped groove small moving ring installed on the impeller cover plate and rotates one by one, and 213 is a V-shaped groove air gap clearance For the two-phase flow inlet, 214 is an inflatable drive two-phase inflow tube that passes through the housing and the impeller cavity cover plate and enters the inlet 213, 215 is the impeller cavity cover plate, 216 is the front end cavity of the inflatable impeller, and 217 is arranged on the impeller to pass through The front and rear cavities in the front cover fixing rivets of the blades communicate with the pressure equalization holes, 218 is an aerated rear cavity, 219 is a casing guide wheel, and 220 is a closed impeller.

In this example, the friction-reducing closed-type constant-velocity high-potential ratio impeller centripetal booster module, the mountain-centred guide wheel 219, the closed-type high-velocity-ratio specific impeller 220, the impeller cavity cover plate 215, and the V-shaped groove air blocking gap annular cover Plates 211, V-shaped groove moving ring 212, two-phase inflow pipe 214, front and rear cavity pressure equalization holes 217 in the front cover rivet, and other internal friction reducing parts or structures are combined. Among them, the rotating curved surface on the guide wheel 219 and the impeller cavity cover 215, and the mating curved surface of the flow path of the outer edge transfer section of the two are specially designed to meet the parameters of the closed-type average speed high-potential ratio impeller 220. The space occupied by the impeller cover is cut out from the cavity cover 215 so that the flow path of the blade groove and the outlet flow path are in abutment. During operation, the relative vortex is blocked by the uniform speed bifurcation, the liquid flow is laminarized at a low speed, and is accelerated in the acceleration section, and the outlet flow velocity is reduced by an equal amount. The ring groove on the impeller cavity cover plate 215, the ring cover plate 211, and the rotation The V-groove air-blocking gap formed by the moving ring 212 can maintain the internal friction reduction state. The pressure equalization hole 217 communicates the back-end cavity with the front-end cavity and reduces the friction of the isobaric filling gas without the need for additional takeover. The inlet tube 214 is connected to a two-phase flow driving pressure source such as a jet, and can drive the double-end cavity to reduce friction.

 The assembly requirements for this module are the same as the example shown in Figure 21. During operation, the impeller output potential ratio can reach 3 ~ 9, and the impeller speed is above 20 meters / second. The combined internal friction reduction technology will reduce the disc friction loss of the closed impeller by 82%-95%, which will increase the pump efficiency by 5-9%. This module is more efficient than all the previous examples. The technological benefits of the modular design and the combined benefits with other modules will make the combined structure of the modules in this example one of the hot selection components in the design of a centrifugal pump.

 Referring to FIG. 26, a centrifugal supercharging module with a modular combination of a uniform speed high potential ratio closed impeller, an internal friction reduction assembly, a pre-rotator, and a centrifugal guide wheel is shown. Among them, 221 is an axial incoming flow pre-rotator, 222 is an air gap structure such as a V-shaped small moving ring that is installed on the impeller cover and rotates one by one, and 223 is an inflation that enters through the casing and the impeller cavity cover. Drive the two-phase inflow tube, 224 is the impeller cover, 225 is the impeller cavity cover, 226 is the front cavity of the inflatable impeller, 227 is the pressure equalizing hole of the front and rear cavity of the impeller arranged in the fixed rivet of the front cover, and 228 is the housing Centripetal guide wheel, 229 is the aerated rear cavity, and 230 is an impeller with an even velocity and a high potential ratio.

In this example, the pre-spinning friction-reducing closed-type constant-velocity high-potential ratio impeller centripetal booster module, the centripetal guide wheel 228, the closed-type average-velocity high-potential ratio impeller 230, the impeller cavity cover plate 225, and the axial incoming flow pre-spinning The device 221 is composed of a V-shaped groove moving ring gas blocking structure 222 and a two-phase inflow pipe 223. Among them, the rotating curved surface on the guide wheel 228 and the impeller cavity cover 225, and the mating curved surface of the flow path of the outer edge transfer section of both are specially designed to meet the parameters of the closed-type average speed high-potential ratio impeller 230. During operation, there is no relative vortex in the blade groove, and the liquid flow will be in the acceleration section of the flow channel. 3

Accelerate, thereby reducing the outlet flow rate by the same amount. The wheel hub is sleeved on the impeller shaft sleeve and rigid ribs. The pre-spinner 221 fixed on the impeller cover is used to load the pre-spin in the axial direction. The V-shaped groove gas blocking structure 111 on the impeller cavity cover plate communicates with the front and rear cavity pressure equalization holes 227 and so on to form an internal friction reducing component. After a two-phase flow driving pressure source such as a jet can drive the dual-end cavity friction reducing.

 The assembly requirements for this module are the same as the example shown in Figure 21. During operation, the closed-type average-velocity high-potential ratio impeller 230 increases the relative flow velocity through the acceleration section at the tail of the flow channel, reduces the outlet flow velocity by the same amount, and greatly improves the diversion efficiency. The combined pre-spinner 221 generates constant velocity pre-spinning in the axial incoming flow, makes the direction of the impeller inlet speed field adaptive to the change of working conditions, improves the impeller efficiency and can avoid cavitation. Combined internal friction reduction technology will increase pump efficiency by 5 to 9%. The impeller output momentum ratio can reach 3 to 9, and the upper limit of the impeller speed can reach more than 20 meters per second. This module is more efficient than all the previous examples. The modular design process benefits and the combined benefits with other modules will make the combined structure of this example module one of the hottest components in the design of a centrifugal pump.

Referring to FIG. 27, a centrifugal supercharging module of a modular combined average speed high potential ratio closed impeller, an ultra-reduction component, a pre-rotator, and a centrifugal guide wheel is shown. Among them, 231 is an axial incoming flow pre-rotator, 232 is a V-shaped groove moving ring installed on the impeller cover plate and rotating together, and 233 is an inflatable drive two-phase inflow pipe entering through the casing and the impeller cavity cover plate. 234 is the impeller cover that extends and contains the cross section of the impeller cavity section of the transfer channel, 235 is the impeller cavity cover, 236 is the uniform speed high potential ratio closed impeller, and 237 is the transfer of the friction reduction after the impeller cover is extended. the impeller chamber portion of the channel cross section of the flow section, the guide wheel 238 from the housing into the rear chamber into a pneumatic tube, the guide wheel 239 is pressurized flow passage ²⁴⁰ is a guide wheel with radial housing.

In this example, the super-anti-friction pre-spinning closed-velocity high-potential ratio impeller centripetal booster module consists of a centrifugal guide wheel 240 and an impeller cover plate 234 with a section of the impeller cavity section that extends and covers the transfer section. The speed high potential ratio impeller 236, the impeller cavity cover plate 235, the axial inflow pre-rotator 231, the super friction reducing component 232, and the two-phase inflow pipe 233 are composed of components. among them, The rotating curved surface on the guide wheel 228 and the impeller cavity cover 225, and the matching curved surface of the flow channel of the outer edge transfer section of the impeller 230 are specially designed to meet the parameters of the closed-type average high-potential ratio impeller 230. During operation, there is no relative vortex in the blade groove, and the liquid flow will be accelerated in the acceleration section of the flow channel, thereby reducing the outlet flow rate by the same amount. The pre-spinner 221 is used to apply a pre-spin to the axial inflow. The V-shaped groove air-blocking gap cover plate installed on the impeller cavity cover plate, the moving ring 232 installed on the impeller cover plate and rotating one way, and the ring groove at the entrance of the impeller cavity cover plate constitute the V-shaped groove air-blocking gap. As an important component of super friction reduction, the front end cavity of the impeller cavity portion extending to the flow channel of the transfer section is not leaked. The rear cavity is individually inflated at a slightly lower pressure. Its inlet tube 238 passes through the shell of the guide wheel and is sealed in. It can be filled with pure gas with controlled flow rate. It can also charge the two-phase flow pressure source of the front cavity such as the ejector through the regulating valve section. After the stream is divided into streams, the pressure drop is adaptive. The idler shaft sleeve may be connected to a static seal or a positive pressure liquid seal, so there is no risk of air leakage.

 The assembly requirements of this module are basically the same as the example shown in Figure 21, and care should be taken to protect the slightly enlarged impeller cover during assembly. During operation, the closed-type average speed high-potential ratio impeller 236 increases the relative flow velocity through the acceleration section at the end of the flow channel, reduces the outlet flow velocity by an equal amount, and greatly improves the diversion efficiency. The combined pre-spinner 231 generates constant-speed pre-spinning in the axial incoming flow, makes the direction of the impeller inlet speed field adaptive to operating conditions, improves the efficiency of the impeller, and avoids cavitation.

Ultra friction reduction technology reduces the disc friction loss by 82% ~ 95% and improves the pump efficiency by 5-9%. The benefits are exactly the same as the internal friction reduction technology. At the same time, it can also make the high-speed friction zone of the transfer section flow path confluent. Most of the losses are eliminated. Calculated according to a uniform circular motion, this part of the loss is proportional to the square of the absolute velocity of the inlet guide, proportional to the length of the manifold and determined by the diameter of the guide wheel and the number of vanes, and proportional to the length of the frictional arc of the cross section. Eliminating this part of the marginal friction actually belongs to the benefit of reducing the diversion loss coefficient. It can be estimated according to the equivalent diameter of the flow channel in the transfer section and its expansion rate, the flow velocity and its change, and the viscosity of the medium, or it can be used to accurately calculate the loss along the way. . These calculations are related to specific flow parameters. The average speed high potential ratio impeller output potential-to-momentum ratio in the module in this example can reach 3 ~ 9, and the impeller speed can be selected above ²⁰ meters / second. The efficiency of this module is higher than all the previous modules, and it is the best example. Its energy-saving benefits, coupled with the modular design process benefits, and the benefits that can be generated by combining with other modules, will make the combined structure of the module of this example the hottest choice component in the centrifugal pump design.

 Figures 28 to 36 are examples of a centrifugal single-stage centrifugal pump constructed by combining a centripetal booster module and a symmetrical end cap module according to a modular approach. The following first describes the common features and advantages of centrifugal single-stage centrifugal pumps in general, and then describes the individual characteristics and effects of each pump in a table.

 According to the modular combination method of the present invention, the centripetal booster single-stage centrifugal pump includes a centripetal booster module, which is one of the example modules shown in FIG. 19 to FIG. 27. The impellers in these modules include different technologies or Process and therefore have different characteristics. The housing of the centripetal booster module is a part of the integrated structure of the centrifugal guide wheel. It has a circular cylindrical shape, a mating surface with a positioning stop and a sealing groove, or a screw through-hole nose structure protruding outward. The front side of the baffle of the guide wheel structure is the impeller cavity. Its shape and position are suitable for installing the impeller with proper clearance. The edge of the baffle has a curved surface that merges with the curved surface of the impeller cavity cover.

 According to the modular combination method of the present invention, the assembly dimensions and basic interface parameters of the module of the same parent specification are the same, and have the compatibility of checking the table to check the interchangeability of the same child specification, and the compatibility of the two assemblyable modules has specifications. . Its interchangeability covers the design process, the production process after design and the use process after production.

Based on these conditions, the specific scheme of the modular combined single-stage centrifugal pump of the present invention is: It includes two variable-angle outlet pipe symmetrical end-cap modules and a centripetal booster module, and the two modules have complete interaction with each other according to corresponding sub-specifications. Replacement, or according to the corresponding parent specifications through a table to check the medium, maximum speed, maximum temperature, maximum pressure and other parameters interchangeability is established, according to "liquid flow from the paraxial annular mouth with the inflow and outflow, the connection mode The 3 modules are axially combined to form a symmetrical cover with modular interchangeability. The combination of booster single-stage centrifugal pump refers to the connection and coordination in the design, the assembly in production and the repair in use, and the interchangeability covers these processes.

 According to the modular combination method, the centripetal booster module and the symmetric end cover module in the solution are also planned through parameters. The former is composed of a radial guide wheel, a closed impeller and an impeller cavity cover, and the latter is Single piece with symmetrical end caps with central worm and kiss tube, 2 symmetrical end caps are used as front end cap and rear end cap respectively. There are also shafting components in the pump, including shafts, bearings, keyways and keys, soft seals made of organic materials, etc., where the shaft is a single-stage standardized part.

 In production, assembly and repair, the centrifugal booster module is assembled in the order of the guide wheel, the impeller and the cover of the impeller cavity, and the three are respectively positioned by the casing stop, the shaft and the impeller cavity of the guide wheel, or when there are other accessories. Opportunistic assembly according to its specific position. The disassembly procedure is the reverse of the assembly process.

 During operation, the liquid flow enters the central volute of the front cover from the inlet pipe at a uniform speed, and is converted into a three-dimensional movement by the wall constraint. It flows from the paraxial annular port through an angular distance ranging from 0 to 360 degrees, and enters the energizing module in the axial direction with a circular amount. The medium-rotating impeller flow channel receives the blade normal force work along the way to accelerate and integrate the centrifugal force work to increase the specific energy, and then enters the guide wheel through the transfer section flow channel, decelerates and pressurizes it, and turns 90 degrees from the near-axis annular exit belt. The loop flows out, then converges into the central volute of the back cover, is constrained in three dimensions by the wall surface, and flows out from the kissing tube through an angular distance ranging from 0 to 360 degrees.

The advantages of the above process are: 1. The loop quantity is conserved throughout the process, the diversion load is light, the time and space change rate is small, the flow field stability is good, and inappropriate "tossing," and less, so the loss is reduced; 2. Between modules The combination of the conservative loop and the shunting and converging process reduces the local loss, avoids the problem of improper constraints before shunting and the under-constraints in shunting in the prior art, and the inlet characteristics of the impeller are better; 3. The confluence and shunt processes are without local excitation, which is the result of the special design of the flow channel of the transfer section of the present invention. In all centrifugal booster centrifugal pumps, except for the non-synchronous pre-spinner, impact loss may occur at the impeller inlet, The remaining types have better process constraints, with most of them Adaptability to changing conditions. In particular, those who have a pre-rotation average speed high potential ratio mechanism are fully constrained throughout, have the highest hydraulic efficiency and complete adaptability to changing conditions. For those with internal friction reduction mechanism, the internal efficiency and total efficiency can be increased by 5-9% alone.

 The common features and advantages of this modular combination of single-stage centrifugal pumps are:

 1. The centrifugal pressurized diversion structure has the smallest volume, the lowest cost, and high efficiency.

 2. The centripetal booster module has the module connection specification for the fluid flow in and out of the paraxial annular port with a bad amount in and out.

 3. Compared with the traditional shell, the symmetrical end cap has a greatly reduced volume, so the cost is lower and the efficiency is higher.

 4. Symmetrical end caps are common to the front and rear of the axially packaged module, and are used in single and multi-stage. Its wormway has the norms of module connection from the paraxial ring-shaped mouth with the amount of axial inflow and outflow, and its outlet pipe has variable angle installation. The adaptability of the snail and the worm produces a linear and slewing motion that transforms efficiently. Its connection function is complete, the manufacturing cost is low, and the application range is wide.

 5. On the whole, the above-mentioned connection specification, module interchangeability, and universality embody or potentially contain the technical and economic benefits of the design, manufacturing, and use process brought by the modular combination method, including reducing workload, shortening the construction period, Increasing convenience, miniaturizing product types and material accessories specifications, reducing regulatory technical barriers, accelerating technology and material transfers, reducing costs by multiple factors, and so on.

 Different technical examples of symmetric cover variable-angle outlet pipe centrifugal booster centrifugal pumps have different characteristics and performance differences. These different personality characteristics and their functional performance characteristics are listed in Table 11 below. Personality Characteristics and Effect Theory of Symmetrical Cover Variable Angle Outlet Concentric Booster Single Stage Pump

Referring to FIG. 28, a centrifugal pump structure of a modular combination of a semi-open impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 241 is the front end cap and its inlet tube, 242 is the shunt center volute on the front end cap, 243 is the impeller cavity cover, and 244 is the impeller flow path. N2004 / 001413

245 is a semi-open impeller, 246 is a guide wheel section of the runner section of the transfer section, 247 is a centripetal guide wheel with a housing, 248 is a guide wheel booster flow path, 249 is a convergence center volute on the rear cover, 250 It is the back cover and its outlet tube.

 This example is a semi-open impeller centrifugal single-stage centrifugal pump with symmetrical cover and variable angle outlet pipe. It includes a semi-open impeller centrifugal booster module and two variable-angle outlet tube symmetrical end cover modules. The impeller 245, the impeller cavity cover 243, and the centrifugal guide wheel 247 are respectively used as the front cover 241 and the rear cover 250, and are axially combined by a shaft system and a fastener.

 Variable angle outlet tube semi-open centrifugal single-stage booster pump is a new type of centrifugal pump. It has valuable characteristics such as the adaptability of the guided flow to the changing operating conditions. It is suitable for the transformation of the barrel of the existing centrifugal pump. The main benefit is to reduce Cost and user-friendly installation, with the potential to increase efficiency. The output of the impeller is more constant than the liquid flow. The impeller speed is generally about 10 meters per second. The impeller speed can be increased when the equivalent diameter of the runner is increased.

 Referring to Fig. 29, a centrifugal pump structure of a modular combination closed impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 251 is the front end cover and its inlet tube, 252 is the shunt center volute on the front end cover, 253 is the impeller cavity cover, 254 is the impeller flow channel, 255 is the closed impeller, and 256 is the guide wheel section of the flow channel section of the transfer section 257 is a centripetal guide wheel with a casing, 258 is a guide wheel booster flow path, 259 is a convergence center volute on the rear cover, and 260 is the rear cover and its outlet pipe.

 This example is a single-stage centrifugal centrifugal pump with centrifugal pressure booster and closed-end centrifugal pump with symmetric cover and variable angle outlet tube. The impeller 255, the impeller cavity cover 253, and the centripetal guide wheel 257 are used as the front cover 251 and the rear cover 260, respectively, and are formed by axially combining the shaft system and the fastener.

The variable angle outlet tube closed centrifugal single-stage pump is a new type of centrifugal pump, which has the characteristics of adaptability to the changing process and operation conditions of the guided flow. It is suitable for the transformation of the barrel of the existing centrifugal pump. The main benefits are reduced cost and Convenient for users to install, and has more than half-open impeller The potential for greater efficiency gains. The output of the impeller is usually more specific than the liquid flow. The impeller speed is generally about 10 meters per second. The impeller speed can be increased when the equivalent diameter of the runner is increased.

 Referring to Fig. 30, a centrifugal pump structure of a modular combination internal friction reducing closed impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 261 is the front end cap and its inlet tube, 262 is the shunt center worm on the front end cap, 263 is the impeller front cavity choke gap, 264 is the two-phase inflow tube, 265 is the impeller cavity cover, and 266 is the front and rear ends in the rivet The cavity communicates with the pressure equalization hole, 267 is a centripetal guide wheel with a shell, 268 is a closed impeller, 269 is a convergence center worm on the back cover, and 270 is the back cover and its outlet pipe.

 This example is a centrifugal single-stage centrifugal pump with centrifugal pressure-enclosing closed impeller with a symmetrical cover and variable angle outlet tube. The closed impeller 268, the impeller cavity cover 265, the centrifugal guide wheel 267 and the air gap 263, the two-phase inflow pipe 264, and the front and rear cavity communication pressure equalization holes 266 are used as the front cover 261 and the rear cover 270, respectively. It is formed by axial combination of shaft system and fastener.

 Variable angle outlet tube anti-friction closed-type centrifugal booster single-stage pump is a new type of centrifugal pump, which has valuable characteristics such as flow-adjustable operating conditions, and the use of internal anti-friction design can independently improve efficiency by 5%. ~ 9%, so it is particularly suitable for the reconstruction of existing centrifugal pumps. The main benefits are three aspects: reducing costs, improving efficiency, and facilitating user installation. The output of the impeller is more constant than the liquid flow, and the impeller speed is generally about 10 meters per second. When the equivalent diameter of the runner is increased, the speed can be increased accordingly.

Referring to FIG. 31, the figure shows a centrifugal pump structure of a modular combination of a half-open type average speed high potential ratio impeller centrifugal booster module and a symmetrical end cover. Among them, 271 is the front end cover and its inlet pipe, and 272 is the front end cover. The vortex of the diverter center, 273 is the impeller cavity cover, 274 is a half-open average velocity high-potential ratio impeller, 275 is the impeller cavity part of the flow section of the transfer section, 276 is the guide wheel part of the flow section of the transfer section, 277 is Centripetal guide wheel with casing, 278 is the supercharged flow path of the guide wheel, 279 is the convergence center volute on the rear cover, 280 is the rear cover and its Out of the pipe.

 This example is a centrifugal centrifugal booster pump with a symmetrical cover and a variable-angle, high-potential, centrifugal, centrifugal booster pump. End cover module, the former is composed of a half-open average speed high potential ratio impeller 274, impeller cavity cover 273, and centripetal guide wheel 277, and the latter is used as a front cover 271 and a rear cover 280, respectively, through a shaft system and a fastener shaft Direction combination.

 The variable angle outlet semi-open type high potential ratio centripetal booster single-stage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three new component technologies: a symmetrical end cap, a high potential ratio impeller and a centrifugal guide wheel. . Its potential ratio is as high as 3 ~ 9, the pressure coefficient is close to the theoretical value, the diversion loss is reduced by an order of magnitude, and the guidance process is adaptively changed under different operating conditions. Its efficiency is greatly improved, and it is also obvious in reducing manufacturing costs and facilitating installation and use by users. Advantage. The impeller speed can be selected at about 20 meters per second, and the speed can be increased when the equivalent diameter of the runner is increased.

Referring to FIG. 32, a centrifugal pump structure of a modular combination closed-type constant-velocity high-potential-ratio impeller centrifugal booster module and a symmetrical end cover is shown. Wherein the distal end cap 281 and into a tube, the front end cap 282 is split center channel scroll, the impeller chamber cover 283, ²⁸⁴ is closed average rate higher than the potential of the impeller, the impeller chamber 285 is part of the flow channel cross section of the transfer , 286 is the guide wheel section of the runner section of the transfer section, 287 is the centripetal guide wheel with the shell, 288 is the guide wheel booster runner, 289 is the convergence center volute on the rear cover, and 290 is the rear cover and its outlet. tube.

 This example is a single-stage centrifugal centrifugal pump with a closed cover and a constant-velocity high-potential ratio impeller. End cover module, the former consists of closed-type high-velocity ratio impeller 284, impeller cavity cover 283, and centripetal guide wheel 287, and the latter is used as front cover 281 and rear cover 290, respectively. Combined.

Variable angle outlet tube closed high potential ratio centripetal booster single-stage pump is a new type of centrifugal pump. It adopts modular combination design method to combine symmetrical end cover, high potential ratio impeller and centripetal guide wheel. Three major new component technologies. Its potential ratio is as high as 3 ~ 9, the pressure coefficient is close to the theoretical value, the diversion loss is reduced by an order of magnitude, and the diversion process is adaptively changed under different operating conditions. Its efficiency is greatly improved and it is better than the half-open type. It is easy to install and use at manufacturing cost. It also has obvious advantages. The impeller speed can be selected at about 20 meters per second, and the speed can be increased when the equivalent diameter of the runner is increased.

 Referring to Fig. 33, the structure of a centrifugal pump with a modular combination of a pre-spin closed, closed-velocity, high-potential ratio impeller, a centrifugal booster module and a symmetrical end cover module is shown. Among them, 291 is the front end cap and its inlet tube, 292 is the shunt center worm on the front end cap, 293 guides the axial flow pre-rotator, 294 is the impeller cavity cover, 295 is the closed average speed high potential ratio impeller, 296 is The guide wheel part of the cross section of the flow channel of the transfer section, 297 is a centripetal guide wheel with a housing, 298 is a guide wheel booster flow path, 299 is a convergence center volute on the rear cover, and 300 is the rear cover and its outlet pipe.

 This example is a centrifugal single-stage centrifugal pump with pre-spinning and closed-type high-potential ratio centrifugal booster with symmetric cover and variable angle outlet tube. Angle exit tube symmetrical end cover module, the former is composed of closed-type uniform velocity high potential ratio impeller 295, pre-rotator 293 installed in the impeller suction chamber, impeller cavity cover 294, and centrifugal guide wheel 297, and the latter is used as the front The cover 291 and the rear cover 300 are axially combined by a shaft system and a fastener.

 The variable angle outlet tube pre-spinning high potential ratio centripetal booster single-stage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three new components: symmetrical end cap, high potential ratio impeller, and centrifugal guide wheel. Technology, and adding pre-spinners to solve impeller inlet characteristics. Its potential ratio is as high as 3 to 9, the pressure coefficient is close to the theoretical value, and the diversion loss is reduced by an order of magnitude. The whole process is adaptively changed under operating conditions, with good anti-cavitation characteristics, its efficiency is greatly improved, and its manufacturing cost is low and it is easy to install and use. . Its impeller speed can reach about 20 meters per second, and it can be increased when the equivalent diameter of the runner is increased.

Referring to FIG. 34, a centrifugal pump structure of a modular combination friction reduction closed-type high-potential ratio impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 301 is the front end cover and its inlet tube, 302 is the shunt center worm on the front end cover, and 303 is the front end cavity to reduce friction and gas. Clearance, 304 is the anti-friction driving two-phase inflow pipe, 305 is the impeller cavity cover, 306 is the pressure equalization hole in the front and rear cavity communication, 307 is the centripetal guide wheel with the shell, 308 is the closed type average speed high potential ratio impeller, 309 It is the convergence center volute on the back cover, and 310 is the back cover and its outlet tube.

 This example is a centrifugal single-stage centrifugal pump with centrifugal pressure-enclosing closed-velocity closed-velocity high-potential ratio impeller centrifugal booster with symmetrical cover and variable angle outlet tube. Angle outlet tube symmetrical end cover module, the former is composed of closed average speed high potential ratio impeller 308, impeller cavity cover 305, centripetal guide wheel 307, and choke gap 303, two-phase inflow tube 304, front and rear cavity pressure equalization holes 306 The latter is used as the front cover 301 and the rear cover 310, respectively, and is formed by axially combining the shaft system and the fastener.

 Variable angle outlet tube anti-friction closed-type high potential ratio centripetal booster single-stage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three new components: symmetrical end cap, high potential ratio impeller and centripetal guide wheel. Technology, and equipped with internal friction reduction technology to solve the problem of wheel friction. Its potential-to-motion ratio is as high as 3 to 9, the pressure coefficient is close to the theoretical value, the diversion loss is reduced by an order of magnitude, and it has the adaptability to the changing process conditions. The mechanical loss is reduced by 82%-95%, the three major bottlenecks that restrict the efficiency of the pump are not present, the efficiency is improved by up to two percentages, and the manufacturing cost is low and the installation and use are convenient. The impeller speed is not restricted by the friction loss of the disc, and it can reach more than 20 meters per second. The larger the equivalent diameter of the flow path, the higher the impeller speed can be selected, so it is especially suitable for high lift occasions.

 Referring to Fig. 35, the figure shows the structure of a centrifugal pump with a modular combination of friction reduction and pre-spinning type average speed high potential ratio impeller centrifugal booster module and symmetrical end cover. Among them, 311 is the front end cover and its inlet tube, 312 is the axial incoming flow pre-rotator, 313 is the front end cavity friction reducing air gap, 314 is the friction reducing driving two-phase inflow tube, 315 is the impeller cavity cover, and 316 is the front and rear ends The cavity communicates with the pressure equalization hole, 317 is a centripetal guide wheel with a casing, 318 is a closed type average speed high potential ratio impeller, 319 is a convergence center worm on the back cover, and 320 is a rear end cover and its outlet pipe.

This example is a centrifugal single-stage centrifugal pump with centrifugal pressure-increasing anti-friction pre-spinning and constant-velocity high-potential centrifugal centrifugal pump with symmetrical cover and variable angle outlet tube And 2 variable-angle outlet tube symmetrical end cover modules, the former has a high potential 001413

It is composed of impeller 318, impeller cavity cover 315, centripetal guide wheel 317, pre-spinner 312 and choke gap 313, two-phase inflow tube 314, and front and rear cavity pressure equalizing holes 316, the latter being used as front cover 311 and rear Cover 320 is formed by axial combination of shafting and fasteners.

 Variable angle outlet tube anti-friction pre-spin closed high potential ratio centripetal booster single-stage pump is a new type of centrifugal pump, which adopts modular combination design method to combine three major end caps, high potential ratio impeller and centripetal guide wheel. New component technology, and equipped with internal friction reduction technology to solve the problem of disc friction, pre-spinner to solve the problem of impeller inlet characteristics. Its potential-to-motion ratio is as high as 3 to 9, the pressure coefficient is close to the theoretical value, and the diversion loss is reduced by an order of magnitude, and it has particularly valuable adaptability to the whole process of changing conditions. The mechanical loss is reduced by 82% to 95%, and the three major bottlenecks that restrict the efficiency of the pump are not present. The efficiency is increased by two percentages. The efficiency characteristic curve is comprehensively raised, the anti-cavitation characteristics are idealized, and the manufacturing cost Low, easy to install and use. The impeller speed is not restricted by the friction loss of the disc, and it can reach more than 20 meters per second. The larger the equivalent diameter of the flow path, the higher the impeller speed can be selected, which is generally suitable for various applications.

 Referring to Fig. 36, the figure shows the structure of a centrifugal pump with a modular combination of ultra-friction, pre-spin-closed, high-potential high-potential ratio impeller centripetal booster module and a symmetrical end cover. Among them, 321 is the front end cover and its inlet tube, 322 is the axial incoming flow pre-rotator, 323 is the air blocking gap, 324 is the front-end cavity antifriction driving two-phase inflow tube, 325 is the impeller cavity cover, and 326 is the extension coating transfer The impeller cover of the segment flow channel, 327 is a closed-type average speed high-potential ratio impeller, 328 is a rear cavity anti-friction driving inlet pipe, 329 is a centripetal guide wheel with a shell, and 330 is a rear cover and its outlet pipe.

This example is a centrifugal supercharged centrifugal centrifugal centrifugal pump with super-friction and pre-spinning closed-speed high-potential ratio impeller. And two variable-angle exit-tube symmetrical end-cap modules, the former has a closed average speed high-potential ratio of impeller cover 326 with an impeller cover 326 extending to cover the flow path of the transfer section, impeller cavity cover 325, centripetal guide wheel 329, The spinner 322 and the air gap 323, the front-end cavity two-phase inflow pipe 324, and the rear-end cavity two-phase inflow pipe 328 are used as the front cover 321 and the rear cover 330, respectively, and are axially combined by the shaft system and the fastener. Made. Variable angle outlet tube super friction reducing pre-spinning high potential ratio centripetal booster single-stage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three symmetrical end caps, high potential ratio impellers, and centrifugal guide wheels. Large new component technology, and equipped with super friction reduction technology to solve the problem of wheel friction and reduce the friction loss of the flow channel in the transfer section, and equipped with a pre-spinner to solve the problem of impeller inlet characteristics. Its potential-to-dynamic ratio is as high as 3 to 9, the pressure coefficient is close to the theoretical value, the diversion loss is reduced by an order of magnitude, and it has particularly valuable adaptability to the whole process of changing conditions. The mechanical loss is reduced by 82% to 95%, and the three major bottlenecks that restrict the efficiency of the pump are not present. The efficiency increase is the largest, the efficiency characteristic curve is comprehensively raised, the anti-cavitation characteristics reach the ideal state, and the manufacturing cost is not high. Easy to install and use. The impeller speed is not restricted by the friction loss of the disc, and the restriction of the conduction speed is also reduced accordingly, so it has the highest upper limit of the impeller speed. The larger the equivalent diameter of the runner, the higher the impeller speed can be selected. The pump is generally suitable for various applications.

 Figure 37 ~ Figure 46 are examples of a centrifugal multi-stage centrifugal pump constructed by combining a centripetal booster module and a symmetrical end cap module according to a modular approach. The following first describes the common features and advantages of these centrifugal multi-stage centrifugal pumps in general, and then describes the individual characteristics and effects of each pump in a list.

 The centrifugal booster multistage centrifugal pump includes multiple centripetal booster modules. The examples shown in Figure 19 to Figure 27 show some possible modules. Different modules have different characteristics. They are in the technical principle, structure, The differences in technology, as well as the radial and axial positioning of the stop, the dynamic fit clearance, and the cross section synthesis of the flow channel section of the transfer section have been described in detail previously. Due to the spatial periodicity of the centrifugal supercharging module on the assembly structure and flow field parameters, this module has the valuable attribute of axially cascading within the winding limits of the general or special constraint structure of the rotating shaft. This is the principle basis and limitation of the modular construction of the multi-stage pump.

The problem of interchangeability of specifications and check specifications of the parent specification for checking interchangeability has been explained earlier, and it is described in combination with a combination example of a single-stage pump, which is also applicable to the combination example of a multi-stage pump. A special problem of multi-stage pumps is that when the number of stages is high and the pressure or head is high, the maximum pressure resistance segment design of the casing is more economical. Then the interchangeability of the specifications of the front and rear end cover modules and the segmented centripetal booster modules should also be established in the corresponding subdomains.

 Based on this, the specific scheme of the modular combined multi-stage centrifugal pump of the present invention is: it includes 2 variable angle outlet pipe symmetrical end cover modules and a plurality of centripetal booster modules with a maximum of 64, and the two modules are based on the corresponding sub-modules. The specifications are completely interchangeable, or according to the corresponding parent specifications, the compatibility of the parameters such as medium, maximum speed, temperature, and maximum pressure is established through a look-up table. Among them, the maximum pressure withstandability or axial segmentation If it is established, according to the connection mode of "fluid flow in and out from the paraxial annular mouth with ring volume," the symmetrical end cover module is divided into front and rear covers, and the centripetal booster module is axially connected in series in order, and all modules are axially combined. That is to say, a symmetrical cover with variable angle outlet tube and centrifugal booster multistage centrifugal pump with modular interchangeability is formed. The combination refers to connection and coordination in design, assembly in production and repair in use, and interchangeability covers these processes.

 According to the modular combination method, the centripetal booster module and the symmetrical end cap are either parameterized. Among them, the former is an axial combination of a centrifugal guide wheel, a closed impeller and an impeller cavity cover, and the latter is a single part of a symmetrical end cap with a central worm and a kiss pipe, and two symmetrical end caps are used as front end covers, respectively. And back cover. There must be shafting components in the pump, including the shaft, bearings, keyways and keys, shaft seals and stuffing boxes, and soft retaining rings. The shaft is a standardized part that contains stage parameters.

 During production or maintenance, the centrifugal booster module is assembled in accordance with the order of the guide wheel, impeller and impeller cavity cover plate. The three are respectively positioned by the casing stop, the shaft and the impeller cavity of the guide wheel, or according to their specific positions when other accessories are available Opportunistic assembly. Disassembly procedures are reversed.

When the multi-stage pump is running, the liquid flow enters the central volute of the front cover from the inlet pipe at a steady flow rate, and is constrained by the wall surface to be converted into a three-dimensional motion to generate a ring volume, and then flows out from the paraxial annular port. This is a front-end boundary flow segment. The liquid flow shunts into the impeller flow channels of the first energizing module, receives the normal force work of the blades to accelerate along the way and integrates the centrifugal force work to increase the specific energy, and then enters the guide wheel through the transfer section flow channel. After decelerating and supercharging, Rotate 90 degrees from the paraxial annular exit with a circular flow. This is the first energization cycle. Except static pressure N2004 / 001413

In addition to the accumulation and transmission of force, the liquid flow enters and exits the second energization module, and the third and nth ones that enter and exit the series in succession ... the flow parameters of the last energization module are all cycles The repetitive nature of the cycle includes the periodicity of the looped flow diversion at the inlet and the looped flow confluence at the exit. These repetitive periods constitute a multi-level energizing process segment. The liquid flow converges from the end-level energizing module into the central volute of the rear cover, and is constrained and transformed by the three-dimensional components constrained by the wall surface to flow out from the kiss-out pipe at a stable flow rate. This is the rear boundary flow section.

 The advantages of the above process are:

 1. Conservative circulation throughout the process, the diversion load of modules at all levels is reduced, the time and space rate of change is reduced, the stability of the flow field is good, and the deceleration advantage of the multi-stage pump reduces the diversion losses at all levels.

 2. The improvement of the condition of shunting and confluence between modules, coupled with the multi-level low-speed advantage, its local loss will be reduced or eliminated, and its adaptability to changing conditions and anti-cavitation characteristics will also be improved due to the speed reduction.

 3. For those who use the pre-rotating average speed high potential ratio impeller module, due to the complete constraints of the whole process, the multi-stage pump will have the highest hydraulic efficiency and complete adaptability to variable working conditions, and its full hydraulic loss belongs to the pure loss mode along the way. Within a certain number of stages, this loss will decrease as the number of stages increases.

The analysis of the loss along the way of the prespin average speed high potential ratio centrifugal booster multistage pump is as follows: Assuming that the flow velocity of the vane groove does not change with the number of stages, the impeller path loss is inversely proportional to the square root of the number of stages. The square root is proportional. Because the stage flow loss is proportional to the third power of the input speed and inversely proportional to the 1.5 power of the number of stages, the total flow loss is inversely proportional to the square root of the number of stages. It is also assumed that the speed of the vane slot is one order of magnitude lower than the speed of the input guide, and the stage impeller loss is about two orders of magnitude smaller than the stage guide loss. Because all types of losses in the whole process include the multiplicative factor of the series, when the normalized ratio of the total loss of the multi-stage pump to the single-stage pump is used, the proportional relationship between the leaf conduction loss can be applied as the weight factor, and the loss of the stage and the number of stages can be used again. Relationship, you can find the multi-stage pump normalized to a single stage The total loss ratio of the pump is shown in Table 13 as an example.

Table 13 The relationship between the loss ratio and the number of stages for a single-stage pump with a guide loss ratio of 2:98 and the constant velocity of a multi-stage pump with a blade groove

Increasing the number of stages can increase efficiency, which is the advantage of multi-stage pumps. However, if the flow velocity of the lobes is not sufficiently low, the minimum point of the total loss ratio than the concave function will correspond to a smaller number of stages, and the potential for improving efficiency will not exceed an order of magnitude.

 4. For those who use internal friction reduction impeller, the improvement of internal efficiency and total efficiency will exceed the range of 5-9% and approach its upper limit, which is compared with single-stage pump. Compared with the multi-stage pump itself, the upper limit of friction reduction and efficiency will be reduced due to the reduction of the impeller speed, and the impeller linear speed and wheel diameter requirements are inversely proportional to the square root of the number of stages. The benefit will not be as significant as a single-stage pump.

 From the perspective of multi-factor technical and economic values such as efficiency and cost, the common characteristics and advantages of modular multi-stage centrifugal pumps are:

 1. The centripetal booster module is interchangeable, and has the normative connection with each other and with the boundary module. The centripetal booster diversion structure has the smallest volume, the lowest cost, and high efficiency.

 2. The symmetrical end cover module has the versatility of front and back, single and multi-level and the interchangeability related to it, has the standardization of connection with the enabling module, and has the convenience of use with variable angle installation, compared with the traditional shell Its volume is greatly reduced, its manufacturing cost is therefore lower, its range of adaptation is wide, and its efficiency is also high.

3. On the whole, based on the above-mentioned connection specification, module interchangeability, and generality, the design scheme directly reflects or potentially contains the design brought by the modular combination method. Technical and economic benefits of planning, manufacturing, and use processes, including reducing workload, shortening construction periods, increasing convenience, simplifying product models and material accessories specifications, reducing regulatory technical barriers, accelerating technology and material transfer, reducing costs by multiple factors, etc. Wait.

 The characteristics and performance differences of different technical examples of centrifugal multi-stage centrifugal pumps with centrifugal booster pumps with symmetrical caps and variable angles are basically the same as those of single-stage pumps. "The column names and contents are also mostly the same, but the key contents are different. For ease of searching and comparison, they are still listed in Table 12 redundantly.

Table 12 Individual characteristics of centrifugal multi-stage centrifugal pump

Referring to FIG. 37, a multi-stage centrifugal pump structure with a modular combination of a semi-open impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 331 is the central volute of the front cover, 332 is the front cover and its inlet tube, 333 is the impeller cavity cover, 334 is a semi-open impeller, 335 is a centripetal guide wheel with a housing, 336 is an impeller flow channel, and 337 is a transfer Section runner section PT / CN2004 / 001413

Minutes, 338 is the supercharged runner of the guide wheel, 339 is the rear cover and its outlet pipe, and 3 ⁴⁰ is the center volute of the rear cover.

 This example is a semi-open impeller centrifugal multi-stage centrifugal pump with symmetrical cover and variable-angle outlet pipe. The former consists of a semi-open impeller 334, an impeller cavity cover 333, and a centripetal guide wheel 335, and the latter is used as a front cover 332 and a rear cover 340, respectively, and is connected and combined through a shaft system and a fastener.

 The variable angle outlet tube semi-open centrifugal booster multistage pump is a new type of centrifugal pump, which has the characteristics of adaptability to the changing process and operation conditions of the guided flow. It is suitable for the simple modification of the existing centrifugal pump. The main benefits are reduced cost and Easy to install and has the potential to improve efficiency. The variable inlet and outlet pipe characteristics can greatly reduce the size of the centrifugal pump. The output of the impeller is usually more specific than the liquid flow. The impeller speed is generally about 10 meters per second. The impeller speed can be increased when the equivalent diameter of the runner is increased. The head of a multi-stage pump is directly proportional to the number of stages. Increasing the number of stages can achieve a high head. For determined head and flow requirements, mathematically planned modules can reduce the total cost of ownership for the user.

 Referring to Fig. 38, a multistage centrifugal pump structure of a modular combination closed impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 341 is the central volute of the front cover, 342 is the front cover and its inlet tube, 343 is the impeller cavity cover, 344 is the closed impeller, 345 is the centripetal guide wheel with the shell, 346 is the impeller flow channel, and 347 is the transfer section The runner section of the runner, 348 is the booster runner, 349 is the rear end cap and its outlet tube, and 350 is the center volute of the rear end cap.

 This example is a centrifugal multi-stage centrifugal centrifugal pump with a closed-end impeller and a centrifugal pump with a symmetrical cover and variable angle outlet tube. It consists of a closed impeller 344, an impeller cavity cover 343, and a centripetal guide wheel 345. The latter is used as a front cover 342 and a rear cover 349, respectively, and is connected and combined by a shaft system and a fastener.

Variable angle outlet tube closed centrifugal booster multi-stage pump is a new type of centrifugal pump with diversion Valuable features such as process adaptability and operating adaptability are suitable for the retrofit of the prior art centrifugal pump. The main benefits are reduced costs and easy installation by the user. At the same time, it has the potential to improve efficiency. The efficiency-enhancing performance is better than the semi-open type. Its variable inlet and outlet pipe characteristics can greatly reduce the type specifications of the centrifugal pump. The output of the impeller is usually more specific than the liquid flow. The impeller speed is generally about 10 meters per second. The impeller speed can be increased when the equivalent diameter of the runner is increased. The head of a multi-stage pump is directly proportional to the number of stages. Increasing the number of stages can achieve a high head. For determined head and flow requirements, mathematically planned modules can reduce the user's total cost of ownership.

 Referring to Figure 39, a multistage centrifugal pump structure with a modular combination antifriction closed impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 351 is the front end cover and its inlet tube and central worm, 352 is the impeller cavity cover, 353 is a closed impeller, 354 is a centripetal guide wheel with a housing, 355 is a front-end cavity V-shaped air gap structure, and 356 is The two-phase inflow pipe, 357 is the closed cover of the impeller front, 358 is the pressure equalization hole in the front and rear cavity communication in the rivet, 359 is the cross section of the flow channel of the transfer section, 360 is the rear cover and its outlet tube and the central worm.

 This example is a centrifugal multi-stage centrifugal centrifugal pump with centrifugal pressure-enclosing closed impeller with symmetrical cover and variable angle outlet tube. The cover module is composed of a closed impeller 353, an impeller cavity cover 3S2, a centripetal guide wheel 354 and a choke gap 355, a two-phase inflow pipe 356, and a front and rear cavity communication pressure equalization hole 358. The latter is used as the front cover 342, respectively. And back cover 349, connected by shafting and fasteners. Among them, the choke gap 355, the two-phase inflow pipe 3S6, and the front and rear cavity communication pressure equalizing holes 358 constitute an intra-stage friction reducing device. In a multi-stage pump structure with no shaft seal in the rear cavity, the driving two-phase inflow pipe is connected from the proximal shaft of the front cavity, and a vent hole is opened at the blade wide or the center of the rivet to keep the pressure in the front and rear cavity equal. Used in some cases to save tubing. Its pressure equalizing effect is similar to the pressure equalizing hole in the prior art, but the pressure equalizing effect is much more significant than that of the liquid phase balancing hole, and it does not cause any volume loss.

Variable angle outlet tube antifriction closed-type centrifugal booster multistage pump is a new type of centrifugal pump with N2004 / 001413

The internal friction reduction thus reduces the internal mechanical loss of 82 to 95%, and the valuable characteristics such as the adaptability of the operating process to changing conditions. It is suitable for the simple modification of the existing centrifugal pump. The main benefits are reduced costs and convenient installation by the user while improving efficiency. Potential, synergistic performance is better than half-open. Its variable inlet and outlet pipe characteristics can greatly reduce the size of the centrifugal pump. The output of the impeller is more constant than the liquid flow. The impeller speed is generally about 10 meters per second. The impeller speed can be increased when the equivalent diameter of the runner is increased. Increasing the number of stages can achieve high heads or improve efficiency. Mathematically planned modules can reduce the total cost of ownership for the user.

 Referring to FIG. 40, a multi-stage centrifugal pump structure with a modular combination of a semi-open type uniform speed high potential ratio impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 361 is the central volute of the front cover, 362 is the front cover and its inlet tube, 363 is the impeller cavity cover, 364 is a half-open average speed high potential ratio impeller, 365 is a centripetal guide wheel with a casing, and 366 is an impeller. 367 is the cross-section guide wheel section of the flow channel section of the transfer section, 368 is the supercharged flow path of the guide wheel, 369 is the rear end cover and its outlet pipe, and 370 is the center volute of the rear end cover.

 This example is a semi-open centrifugal multi-stage centrifugal pump with centrifugal booster pump with symmetrical cover and variable angle outlet pipe. Tube symmetrical end cover module, the former is composed of a half-open average speed high potential ratio impeller 364, impeller cavity cover 363 and centripetal guide wheel 365, and the latter is used as the front cover 362 and the rear cover 369, respectively, through the shaft system and fastening Piece connection combination.

The variable angle outlet semi-open type high potential ratio centripetal booster multi-stage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three new component technologies: a symmetrical end cap, a high potential ratio impeller and a centrifugal guide wheel. . Its potential ratio is as high as 3 ~ 9, the pressure coefficient of the stage is close to the theoretical value, the stage diversion loss is reduced by an order of magnitude, and the guidance process is adaptively changed under different operating conditions. Its efficiency is greatly improved, and it reduces manufacturing costs and facilitates installation and use by users. It also has obvious advantages. Its variable inlet and outlet pipe features are convenient for users to install, and can also greatly reduce the type specifications of the centrifugal pump. The impeller speed can reach about 20 meters per second, and the impeller speed can be increased when the equivalent diameter of the runner is increased. Increase the number of stages of a multistage pump Can reach] to a very high head, or further improve efficiency. Mathematically planned modules can reduce the total cost of ownership for users.

 Referring to FIG. 41, a multi-stage centrifugal pump structure with a modular combination closed-type high-potential ratio impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 371 is the central volute of the front cover, 372 is the front cover and its inlet tube, 373 is the impeller cavity cover, 374 is a closed-type average speed high potential ratio impeller, 375 is a centripetal guide wheel with a casing, and 376 is an impeller flow channel. 377 is the guide wheel section of the runner section of the transfer section, 378 is the supercharged runner of the guide wheel, 379 is the rear end cover and its outlet pipe, and 380 is the volute in the rear end cover.

 This example is a symmetric cover variable angle outlet tube closed type average speed high potential ratio impeller centrifugal booster multi-stage centrifugal pump, which contains a plurality of closed type average speed high potential ratio impeller centrifugal booster modules and 2 Variable-angle outlet tube symmetrical end cover modules. The former consists of a closed-type high-speed ratio impeller 374, an impeller cavity cover 373, and a centripetal guide wheel 375. The latter is used as a front cover 372 and a rear cover 379, respectively. And fastener connection combinations.

 The variable angle outlet pipe closed high potential ratio centripetal booster multistage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three major new component technologies: a symmetrical end cap, a high potential ratio impeller, and a centrifugal guide wheel. Its potential ratio is as high as 3 to 9, the stage pressure coefficient is close to the theoretical value, the stage diversion loss is reduced by an order of magnitude, and the guidance process is adaptively changed to operate under conditions. The efficiency is greatly improved, and the efficiency is due to the half-open type. It also has obvious advantages in reducing manufacturing costs and facilitating user installation and use. Its variable inlet and outlet pipe features are convenient for users to install, and can also greatly reduce the size of the centrifugal pump. The impeller speed can reach about 20 meters per second, and the impeller speed can be increased when the equivalent diameter of the runner is increased. Increasing the number of stages of a multi-stage pump can achieve high heads or further increase efficiency. Mathematically planned modules can reduce the total cost of ownership for users.

Referring to FIG. 42, a multi-stage centrifugal pump structure with a modular combination of a pre-spinning closed-velocity high-potential ratio impeller centrifugal booster module and a symmetrical end cover is shown. Among them, 381 is the central volute of the front cover, 382 is the front cover and its inlet tube, 383 is the impeller cavity cover, 384 is a closed-type average speed high potential ratio impeller, 385 is a centripetal guide wheel with a casing, and 386 is a prespin , 387 Is the impeller acceleration flow path, 388 is the guide wheel booster flow path, 389 is the rear end cover and its outlet pipe, and 390 is the center volute of the rear end cover.

 This example is a centrifugal multistage centrifugal pump with centrifugal pump with pre-spinning closed-velocity and high-potential ratio impeller. The module consists of a closed-end high-potential ratio impeller 384, a pre-rotator 386, an impeller cavity cover 383 and a centripetal guide wheel 385, and two variable-angle outlet symmetrical end-cap modules, and the latter is used as the front cover 382, respectively. It is combined with the rear cover 389 through a shaft system and fasteners.

 The variable angle outlet tube closed high potential ratio centrifugal booster multistage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three major new component technologies: a symmetrical end cap, a high potential ratio impeller, and a centrifugal guide wheel. Its potential-to-moment ratio is as high as 3 to 9, the stage pressure coefficient is close to the theoretical value, and the stage diversion loss is reduced by an order of magnitude. Its pre-spinner completely eliminates the impact of turbulence and cavitation in the impeller inlet area, enables the entire process to adaptively change operating conditions, and greatly improves its efficiency. It also has obvious advantages in reducing manufacturing costs and facilitating user installation and use. Its variable inlet and outlet pipe features are convenient for users to install, and can also greatly reduce the size of the centrifugal pump. The impeller speed can reach about 20 meters per second, and the impeller speed can be increased when the equivalent diameter of the runner is increased. Increasing the number of stages of a multi-stage pump can achieve very high heads or further increase efficiency. Mathematically planned modules can reduce the total cost of ownership for users.

Referring to FIG. 43, the figure shows a multi-stage centrifugal pump structure with a modular combination of friction reducing closed-type high-potential ratio impeller centrifugal booster module and symmetrical end cover. Among them, 391 is the central volute of the front cover, 392 is the front cover and its inlet tube, 393 is the impeller cavity cover, 39 ⁴ is the closed type: 4 high potential ratio impeller, 395 is a centripetal guide wheel with a housing, and 396 is the impeller Front-end cavity V-shaped groove air-blocking gap, 397 is the friction-reducing driving two-phase inflow tube, 398 is the pressure equalization hole in the front and rear cavity of the impeller cover fixing rivet, 399 is the rear end cover and its outlet tube, 400 is the rear end cover Central worm.

This example is a centrifugal centrifugal pump with centrifugal booster and closed-velocity closed-velocity high-potential ratio impeller with symmetric cover and variable angle outlet tube. Centripetal booster module and two variable-angle outlet tube symmetrical end cover modules, the former consists of closed-type high-speed ratio potential impeller 394, impeller cavity cover 393, centripetal guide wheel 395 and choke gap 396, and two-phase inflow pipe 397 The front and rear chambers are connected by pressure equalization holes 398, which are used as the front cover 392 and the rear cover 399, respectively, and are connected and combined by a shaft system and a fastener. Among them, the air blocking gap 396, the two-phase inflow pipe 397, and the front and rear chambers communicating with the pressure equalization hole 398 constitute an intra-stage friction reducing device. This structure saves pipelines without any volume loss.

 Variable angle outlet tube anti-friction closed-type high potential ratio centrifugal booster multistage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three new components: symmetrical end cap, high potential ratio impeller, and centrifugal guide wheel. technology. Its potential-to-moment ratio is as high as 3 to 9, the pressure coefficient of the stage is close to the theoretical value, the stage diversion loss is reduced by an order of magnitude, and the guidance process is adaptively operated under variable operating conditions, which greatly improves the efficiency. The combined internal anti-friction device can reduce wheel friction loss by 82% ~ 95%, which can further improve efficiency. This pump also has obvious advantages in reducing manufacturing costs and facilitating user installation. The variable inlet and outlet pipe angles are convenient for users to install, and the type specifications of centrifugal pumps can be greatly reduced. The impeller speed can reach about 20 meters per second, and the impeller speed can be increased when the equivalent diameter of the runner is increased. Increasing the number of stages of a multi-stage pump can achieve very high heads or further increase efficiency. Mathematically planned modules can reduce the total cost of ownership for users.

Referring to FIG. 44, the figure shows a multi-stage centrifugal pump structure with a modular combination anti-friction pre-spin-closed average speed high potential ratio impeller centrifugal booster module and a symmetrical end cover. Among them, 401 is the pre-rotator suspension rib, 402 is the front end cover and its inlet tube and central worm, 403 is the impeller cavity cover, 404 is the pre-spinner, 405 is the centripetal guide wheel with the shell, and 406 is the closed type average speed higher potential than the wheel, anti-friction drive 407 is a two-phase inlet pipe, the air impeller ⁴⁰⁸ is distal chamber barrier gap, the impeller 409 is a head end chamber in communication with the front and rear fixing rivet pressure equalizing hole and the rear cover 410 is a tube And the center worm.

This example is a centrifugal multi-stage centrifugal centrifugal pump with centrifugal pump with symmetrical cover and variable angle outlet for reducing friction and pre-spinning high speed potential. Concentric pressurization module and 2 variable angle outlet tube symmetrical end cover modules, the former consists of 004 001413

The closed-type average speed high potential ratio impeller 406, pre-rotator 404, impeller cavity cover 403, centripetal guide wheel 405 and choke gap 408, two-phase inflow pipe 407, front and rear cavity communication pressure equalization holes 409, the latter are respectively Used as the front cover 402 and the rear cover 410, and connected by a shaft system and a fastener. Among them, the air blocking gap 408, the two-phase inflow pipe 407, and the pressure equalization hole 409 constitute a friction reducing device in the module.

 Variable angle outlet tube anti-friction pre-spin closed high potential ratio centrifugal booster multistage pump is a new type of centrifugal pump. It adopts modular combination design method to combine three major end caps, high potential ratio impeller and centripetal guide wheel. New component technology. Its potential momentum ratio is as high as 3 to 9, the stage pressure coefficient is close to the theoretical value, and the stage conduction loss is reduced by an order of magnitude. The combined anti-friction device can reduce the friction loss of the disc by 82% ~ 95%, which can further improve the efficiency. The combined pre-spinner completely eliminates the impact of turbulence and cavitation in the impeller inlet area, making the whole process adaptively change the operating conditions. This pump also has obvious advantages in reducing manufacturing costs and facilitating installation and use by users. The variable inlet and outlet pipe angles are convenient for users to install, and the type specifications of centrifugal pumps can be greatly reduced. The impeller speed can reach about 20 meters per second, and the impeller speed can be increased when the equivalent diameter of the runner is increased. Increasing the number of stages of a multi-stage pump can achieve very high heads or further increase efficiency. Mathematically planned modules can reduce the total cost of ownership for users.

 Referring to FIG. 45, the figure shows a multi-stage centrifugal pump structure with a modular combination of ultra-friction, pre-spin closed, high-potential average speed high-potential ratio impeller centrifugal booster module and symmetrical end cover. Among them, 411 is a pre-rotator, 412 is a front end cover and its inlet tube and a central volute, 413 is an impeller cavity cover, 414 is an average speed high potential ratio closed impeller, 415 is a centripetal guide wheel with a casing, and 416 is an impeller Front-end cavity air-blocking gap, 417 is the anti-friction driving two-phase inflow tube, 418 is the impeller cover covering the impeller cavity part of the runner section of the transfer section, 419 is the rear-end cavity anti-friction driving medium inlet tube, and 420 is the rear end cover With its outlet tube and central worm.

This example is a centrifugal pump with centrifugal pump with centrifugal pump and centrifugal pump for centrifugal superimposed friction and closed-velocity high-potential ratio centrifugal pumps. Centripetal booster module and 2 variable angle outlet symmetrical end sense modules, the former consists of Impeller cover 418 with extended and covered transfer section closed closed average speed high potential ratio impeller 414, pre-spinner 411, impeller cavity cover 413, centripetal guide wheel 415 and air gap 416, front end cavity anti-friction drive The two-phase inflow pipe 417 and the rear cavity anti-friction driving medium inlet pipe 419 are composed of the latter used as the front cover 412 and the rear cover 420, respectively, and connected and combined by a shaft system and a fastener. Among them, the gas blocking gap 416 and the two-phase inflow pipe 417 are front-end cavity super-friction driving components, and the rear-end cavity anti-friction driving medium inlet pipe 419 is separately input to the two-phase flow or gas-driven rear-end cavity anti-friction.

 The variable angle outlet tube super friction reducing pre-spinning high potential ratio centripetal booster multistage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three symmetrical end caps, high potential ratio impellers, and centrifugal guide wheels. The large new component technology has a potential-to-moment ratio of 3 to 9, the pressure coefficient of the stage is close to the theoretical value, and the stage conduction loss is reduced by an order of magnitude. Equipped with super friction reduction technology to solve the friction problem of the disk and reduce the friction loss of the flow channel in the transfer section, reduce the friction loss of the disk by 82% ~ 95%, greatly reduce the flow channel loss in the transfer section, and further improve the efficiency. The combined pre-spinner completely eliminates the impact of turbulence and cavitation in the impeller inlet area, and enables the entire process to adapt to variable operating conditions. This pump also has obvious advantages in reducing manufacturing costs and facilitating installation and use by users. Its variable inlet and outlet pipe characteristics are convenient for users to install, and it can also greatly reduce the type specifications of centrifugal pumps. The impeller speed can reach about 20 meters per second, and the impeller speed can be increased when the equivalent diameter of the runner is increased. Increasing the number of stages of a multi-stage pump can achieve very high heads or further increase efficiency. Mathematically planned modules can reduce the total cost of ownership for users.

 Referring to FIG. 46, a multi-stage centrifugal pump structure of a modular combination of pre-spinning double semi-open centripetal booster module and symmetrical end cover module is shown. Among them, 421 is the central volute of the front cover, 422 is the front end cover and its inlet tube, 423 is the impeller cavity cover, 424 is a semi-open type average speed high potential ratio impeller, 425 is a semi-open centripetal guide wheel with a housing, 426 Is the impeller flow path acceleration section, 427 is the centrifugal guide wheel pressurized flow path, 428 is the radial inflow pre-rotator, 429 is the rear end cover and its outlet pipe, and 430 is the rotating curved steering shaft sleeve.

This example is a centrifugal multi-stage centrifugal pump with centrifugal booster and centrifugal pump. Contains a maximum of 64 pre-spinning double-half-type average speed high-potential ratio impeller centrifugal booster modules and two variable angle outlet tube symmetrical end cover modules. The former consists of half-open type average-speed high-potential ratio impellers. 424, the impeller cavity cover 423, the semi-open centripetal guide wheel 425 and the pre-spinner 428, the latter being used as the front cover 422 and the rear cover 429, respectively, and connected and combined by a shaft system and a fastener. Among them, the pre-spinner can be axially flowing and installed in the impeller suction chamber, or it can be changed to radial-flowing pre-spinner and installed in the center of the guide wheel (the first stage is not changed, or default, the last stage can be missing province).

 The variable angle outlet pipe pre-spin silent semi-open centrifugal booster multi-stage pump uses a pre-spin average speed high potential ratio double semi-open wheel module as an enabling module. The semi-open guide wheel of this module has only a ring-shaped back cover, which is integrated with the circular cylindrical shell. The blades are fastened to the back cover or manufactured integrally. The back cover is opened to form a diameter and impeller. The cavity with the same inlet is used as the space for passing the shaft, sucking in and installing the pre-spinner. The guide wheel and its shell are concentrically installed with the rotating shaft. The front cover plate shares the bottom surface of the impeller back cover disk. The radial inflow pre-rotator has its own sleeve, which fits tightly on the rotating shaft and is located in the cavity position in the guide wheel. The semi-open impeller only has a single-sided back cover wheel disc, and the blades are fastened to it or are manufactured integrally. The impeller is fastened to the shaft through a shaft sleeve. The impeller cavity front cover is positioned and installed on the front side of the impeller cavity through the guide wheel housing. Double semi-open centripetal booster modules are assembled in series step by step. The pre-spinner is not installed in the last stage. With the front cover and back cover modules, the assembly is complete.

 The above-mentioned half-open multi-stage solution is a simple structure of a multi-stage centrifugal pump with a uniform speed and high potential ratio. The impeller flow channel and the guide wheel flow channel share the impeller back cover as a rotating separation structure, and the guide wheel flow channel and the next stage impeller flow channel share the guide wheel back cover as a static separation structure, and the axial size of the pump is reduced. However, it does not produce significant commutation local resistance loss.

The impeller of the silent semi-open multistage pump has no end cavity, and no friction reducing mechanism is required. Due to the peripheral velocity component of the guide runner flow, the disc friction speed of the rear cover of the impeller is reduced by about half, and its friction loss can be reduced by about 75%. In addition, this friction also generates a momentum moment increase for the liquid flow in the diversion channel, thereby generating specific work outside the dust impeller. impeller N2004 / 001413

The disk friction of the front cover is the same as that of the prior art. Considering the reduction of the rear cover friction and the benefits of the external specific work, the total disk friction loss is reduced by about 40% compared with the prior art.

 Variable angle outlet pipe pre-spinning double semi-open centrifugal booster multistage pump is a new type of centrifugal pump. It adopts a modular combination design method to combine three new parts technology: symmetrical end cap, high potential ratio impeller, and centrifugal guide wheel. . Its potential-to-motion ratio is as high as 3 to 9, the stage pressure coefficient is close to the theoretical value, and the stage conduction loss is reduced by an order of magnitude. The combined pre-spinner completely eliminates the impact of turbulence and cavitation in the impeller inlet area, making the whole process adaptively change the operating conditions, and its efficiency is greatly improved. This pump also has obvious advantages in reducing manufacturing costs and facilitating installation and use by users. Its variable inlet and outlet pipe features are convenient for users to install, and can also greatly reduce the size of the centrifugal pump. The impeller speed can reach about 20 meters per second, and the impeller speed can be increased when the equivalent diameter of the runner is increased. Increasing the number of stages of a multi-stage pump can achieve very high heads or further increase efficiency.

 The main advantages of the variable angle outlet pipe pre-spinning double semi-open centrifugal booster multistage pump are the simple structure and low manufacturing cost. Both the impeller and the guide wheel are manufactured by a two-clamp molding process, and the mold and processing costs are relatively low. Saving the impeller front cover and reducing the axial size are important factors in reducing costs. Compared with other schemes of the present invention, the efficiency of the silent half-open scheme is reduced by about 2-3%, but its cost is the lowest.

 [Definition]

Absolute speed—the speed of the fluid particle in the centrifugal pump relative to the stationary casing. Relative speed-the speed of the fluid particle in the centrifugal pump relative to the rotating impeller. Implication Speed The speed at which the impeller that drives the fluid in a centrifugal pump moves relative to the stationary casing.

Synchronous speed—the flow velocity of the centrifugal pump flow path with the circumferential division equal to the implication speed at the same radial coordinate.

Specific work—the work done by a centrifugal pump impeller on a unit of mass of fluid flowing through the bucket. Specific energy-the mechanical energy of a unit mass of fluid, distinguishing potential energy from kinetic energy PT / CN2004 / 001413

Called specific potential energy and specific kinetic energy.

Potential lift—the ratio of the pressure increase obtained by the fluid in the impeller flow path to the acceleration of gravity. The pressure increase is equal to the algebraic sum of the centrifugal work of force per unit mass of fluid and its relative kinetic energy reduction.

Dynamic lift—the ratio of the kinetic energy gain per unit mass of fluid in the impeller flow path to the acceleration rate of gravity. This increment completes the conversion of the head in the deflector and is calculated according to the absolute speed in the stationary coordinate system.

Potential-Motion Ratio—The ratio of the impeller output potential head to the output dynamic head, or the ratio of specific potential energy increase to specific kinetic energy increase, is referred to as potential ratio.

Reaction degree—the ratio of the impeller output specific potential energy to the specific power of the transmission, which is a monotonically increasing function with the potential dynamic ratio.

Pressure coefficient—The ratio of the effective specific work delivered by the impeller to the specific kinetic energy calculated based on the peripheral speed of the impeller.

Specific number of revolutions—the rotational speed of a geometrically similar unit impeller conveying unit specific work and unit volume flow, also known as similarity coefficient.

Relative vortex—the reverse vortex motion of the fluid in the flow path of the centrifugal pump impeller relative to the circumferential coordinate system of the rotating coordinate system is a fluid inertial motion under the constraint of weak blade shear, which has a significant impact on the distribution of the relative velocity field .

Local excitation—Incompressible fluid is subjected to a small impulse on the active surface in the unconstrained direction, and non-potential flow movement occurs in the vicinity of the active surface, such as stone hitting water, waterfall flow, agitation, and steep expansion of liquid flow cross section.

Complete Constraints—Fluid constraints without the possibility of local excitation, for example, incompressible fluids with free or weakly constrained boundaries, their possible active surfaces and forward neighborhoods are surrounded by rigid walls orthogonal to the active surface.

Coanda effect—When pressure and wall curvature are appropriate, an incompressible fluid with a certain flow rate adheres to the wall surface, and its non-draining condition is that the absolute pressure on the wall surface is greater than the saturated air pressure. High potential ratio impeller——a centrifugal pump impeller with high potential ratio output, its flow path is complete Constrained, it has the characteristics of low flow rate, anti-relative vortex, no backflow and outflow, constant pressure and constant pressure on the circular cylinder, and linear throttling of kinetic energy feedback.

Feedback reduction ratio—the ratio of the relative velocity of the impeller in the tangential outlet of the high potential ratio to the peripheral speed of the impeller, also known as the feedback reduction coefficient. The specific work of the impeller and the absolute speed of the impeller are both linear reduction functions.

 Adaptive pre-spinner Coaxial pre-spinner equipped with variable angle of attack elastic flow channel for incoming pre-spin and cascade velocity field finishing.

Uniform speed bifurcation—a kind of bifurcation structure of impeller with high potential ratio impeller, which mainly distributes the resistance of the vortex mainly through the uneven distribution of the area of the bifurcation exit area, so as to make the velocity distribution of the trough uniform and low-speed laminarization.

Internal friction reduction——A method and design for improving the mechanical efficiency in a centrifugal pump. Dynamically filling the gas phase medium between the impeller end face and the cavity wall, its viscosity coefficient and specific friction are reduced by two orders of magnitude compared with liquid medium. Mount.

Conservative loop design—a centrifugal pump flow field design that conserves liquid flow moment and inertia. It is required to maintain the continuity of the velocity loop in the impeller connection flow path, including synchronous positive pre-spin and axial surface reversal with loop input guide. .

Centripetal guide wheel—an inward scroll type deflector with conservative loop design. The booster flow path is fully constrained, and the cross-sectional area is gradually increased and the radius of curvature of the center line is gradually reduced according to the optimal expansion rate. It has small size and efficiency. High characteristics.

Centripetal pressurization—The pressure distribution characteristic of the centripetal guide wheel, whose flow path pressure monotonously increases as the centerline pole diameter decreases.

Transfer section flow channel—The liquid flow channel between the impeller outlet and the centrifugal guide wheel pressurized flow channel is surrounded by a curved surface outside the impeller cavity wall.

Ultra Friction Reduction——A method and design for improving the efficiency of centrifugal centrifugal centrifugal pumps with internal friction reduction. The front cover of the impeller is extended to include the impeller cavity part of the flow passage in the transfer section. Extend to that part.

Symmetric end cap-a universal front and rear, single and multi-stage universal end cap for the centrifugal pump The bearing pedestal, the sub-convergence center volute, and its kissing pipeline, the front and rear end caps can be rotated to change the inlet and outlet angles during assembly.

Centrifugal booster module-a centrifugal pump energizing booster unit mainly composed of an axial combination of an impeller and a centrifugal guide wheel, with standardized interface parameters and assembly dimensions, single and multi-stage universal, its interchangeability covers design, production And use process.

Modular combination-a method of combining single-stage and multi-stage centrifugal pumps with a centripetal booster module and a symmetrical end cover module across models. , Connection mode.

Quadratic worm trajectory-a centrifugal pump worm trajectory made up of two cross sections of a fixed-length semi-ellipse and a fixed-chord long bow. The cross-sectional area is a quadratic function of the center angle, which can optimize the friction surface and The flow field gradient, the loss is small.

Trapezoidal groove guide ring—a centrifugal pump guide ring with an isosceles trapezoidal cross-section, used as a transitional deflector between the impeller and the volute, to avoid local excitation. The width of the inlet and outlet is equal to the width of the impeller outlet and worm inlet.

Claims

Rights request

1. A centrifugal pump composed of an impeller, a deflector, a casing, and a shaft system component, characterized in that the tail of the impeller's channel of the impeller is bent in the tangential direction and the cross-sectional area is gradually reduced, and the path of the fluid doing work under centrifugal force The end is accelerated and changed direction, and finally flows out of the impeller at a relatively large speed and an exit angle close to 0. The absolute speed of the outlet is correspondingly reduced. The reaction torque generated by the steering and acceleration processes reduces the work of the shaft.

 2. —A kind of centrifugal pump, consisting of impeller, deflector, casing and shafting components, which is characterized by: the front and rear cavity of the closed impeller disc or the rear cavity of the semi-open impeller is placed in the gas circulation or gas In the liquid two-phase flow circulation process, the end cavity is filled with insoluble gas, the impeller disk rotates in the gas phase medium, and the end cavity pressure dynamically maintains a balance with the surface pressure of the rotating liquid flow along the edge of the end cavity during the circulation, and is equal to or less than the impeller outlet Static pressure.

 3. A centrifugal pump consisting of an impeller, a deflector, a casing, and a shaft system component, characterized in that: the deflector is a centripetal guide wheel, the flow channel is inwardly vortex, the radius of curvature is gradually reduced and cut off The area gradually expands, converges in the central ring cavity and rotates 90 degrees to output axially.

 4. A modular combination method of centrifugal pumps. Centrifugal pumps of different models and specifications use energized components of the same specifications, which are characterized by: based on the periodicity of the space of the centrifugal guide wheel, so as to construct a paraxial annular port with a ring quantity. The axially connected centrifugal booster module of the connected impeller is based on the structural characteristics of the symmetric end cap, and it is used as an end seal structural module with a ring-shaped paraxial ring mouth connection. In the corresponding specifications of the two modules, Modules of the same parent specification have the same assembly dimensions and basic interface parameters and have a look-up table to check interchangeability. Modules of the same seed specification under the parent specification have the same assembly dimensions and all interface parameters and are completely interchangeable. Two interchangeability Defined on a large set of single-stage pumps, multi-stage pumps, various models and different built-in technologies, including centrifugal pumps using high-potential or constant-potential impellers, defining interchangeability domains in the planning and design process, and designing The interchangeability during the subsequent production process and in the use process after production is established within the definition domain.

The connection mode of "fluid flow in and out from the paraxial annular mouth with ring volume" will be 1 or Up to 64 serially connected multiple centripetal booster modules combined with 2 symmetrical end cover modules form a modular combination single-stage or multi-stage pump.

 5. The centrifugal pump according to claim 1, characterized in that: the impeller flow channel outlet is rectangular or circular, the angular distance between adjacent outlets is equal to 360 degrees divided by the number of flow channels, and the normal surface of the outlet is perpendicular to the flow channel The connection between the inner side of the previous exit and the outer side of the next exit is a smooth involute arc cylinder surface or a groove surface from deep to shallow, and the single surface between the cylinder surface or the groove surface and the circumferential cylinder surface of the impeller The cross-sectional area of the side-constrained flow channel has a periodic linear relationship with the center angle. The separated exit stream bundles are arranged by the Coanda effect of the arc cylinder or the groove surface, and an inwardly curved uniform distribution is formed in the exit interval of the wheel edge. The radial and tangential components of the streamline and velocity do not change with the center angle. The sum of the outlet area of each flow channel is equal to the ratio of the design volume flow rate and the design outlet relative speed, which is equal to the product of the impeller peripheral speed and the feedback reduction ratio K.

 6. The centrifugal pump according to claim 5, characterized in that: the impeller blades are L-shaped, the front and middle portions are respectively straight line segments, and they move in a radial direction, and the elbows and the tails are bent to the anti-tangential direction with a proper curvature radius transition, The tail has mechanical strength and sharp ends to isolate the internal and external pressure difference. The proper curvature radius transition includes the shaping changes on both the inside and outside. , The outer side of the tail meets the requirements for the change in strike angle of the Coanda effect finishing, and the outer radius of curvature of the elbow meets the condition of no flow.

7. The centrifugal pump according to claim 1 or 5 or 6, characterized in that: the impeller suction chamber or the outlet of the fore-stage deflector is provided with an axial or radial inflow adaptive prespin that rotates coaxially with the impeller. The pre-spinner is composed of elastic sail blades, rims and rigid ribs. The number of blades is less than the number of blades of the impeller. The blades are made of composite material. It has a tensile elastic coefficient that gradually increases from the front end to the root. It is fixed at the assembly position with equal angular distribution of the wheel hub, the wheel hub is freely sleeved on the rotating shaft or the impeller sleeve, the front end of the blade is suspended on a rigid rib at the entrance, and a pre-spinning flow channel is formed between the blades, wherein, the axial direction The rigid ribs of the inflow pre-rotator are arranged in a radial position on the circumferential surface of the inlet. The rigid ribs of the forward flow pre-rotator are arranged at a position parallel to the rotation axis on the entrance cylindrical surface

8. The centrifugal pump according to claim 1 or 5 or 6, characterized in that: a uniform-speed bifurcation to curb relative vortex is arranged in the impeller impeller, and each of the impeller flow channels is combed by 1 to 3 blades of uniform velocity. It is divided longitudinally to form 2 to 4 branch roads. The cross-sectional area of the branch road entrance is close to the entrance of the leaf trough, and its cross-sectional area is uniformly distributed. Distribution and given speed distribution, or experimentally optimized face data distribution.

 9. The centrifugal pump according to claim 2, comprising: a jet device for inflating the friction reducing end cavity, and the driving pressure liquid of the jet device is divided by the outlet of the pump, and its ejection port is connected to the air source through a regulating valve or Through the atmosphere, its outlet is connected to the anti-friction end cavity from the stationary wall near the axis. The two-phase flow is separated in the cavity. The gas is confined in the cavity by the centripetal buoyancy of the centrifugal force field. The liquid and excess gas pass from the side clearance of the wheel. Drain into the deflector, change the leakage gap between the front end cavity and the suction chamber to a gas blocking gap, or install a gas blocking V-ring groove that separates the split two-phase flow, and install a two-phase lubricating soft retaining ring, or directly The pressure liquid is plugged, and the cavity wall generatrix at the impeller outlet is also made into a shape with ejective decompression effect to reduce the end cavity pressure to below the outlet static pressure.

 10. The centrifugal pump according to claim 2, characterized in that: after the pressure of the canned gas is used to reduce the pressure of the pressure reducing valve and the regulating valve to adjust the flow rate, the pipeline is connected to the static wall of the anti-friction end cavity for injection, and is injected from the pump. A small flow of liquid is split at the outlet, and the pipeline is directly injected into the leakage gap of the mechanical seal cavity and the front end cavity to cool the shaft seal and seal the leakage gap, respectively. Alternatively, the pressure canned gas is depressurized by a pressure reducing valve and the flow is adjusted by a regulating valve. The airflow is directly injected into the return pipe leading from the outlet of the pump to form a two-phase flow, which is connected to the rear wall static wall surface and the front cavity static wall surface near the axial gas-blocking gap, respectively. The gas and liquid flows are adjusted separately. The flow regulating valve is connected in series in the return pipe leading from the pump outlet.

11. The centrifugal pump according to claim 3, wherein: the centrifugal guide runner The transfer section is surrounded by the impeller exit cylinder, the curved surface of the front wall of the impeller cavity and the curved surface of the front bottom surface of the guide wheel. The section is divided into the impeller cavity part and the guide wheel part. The two parts are assembled and joined together. The position of the composite section is periodic. The direction of the ground guide wheel is shifted, and its cross-sectional area increases linearly periodically with the increase of the diversion center angle. Its period is equal to the center angle of a diversion channel, and its increase proportionality factor is equal to that of the impeller rotating through a unit angle. The ratio of the design value of the volumetric displacement to the design value of the absolute speed of the liquid flow outlet, or multiply it by an expansion coefficient greater than 1 and smaller than the minimum expansion rate of the booster runner of the guide wheel.

 12. The centrifugal pump according to claim 3, characterized in that: a symmetrical end cap with a central vortex confluent angle outlet pipe is used for front and rear axial packaging, and the end cap is composed of a pressure-bearing cover plate and a cap with an assembly stop. The integrated structure of the plate is composed of a central structure and a connecting pipeline. The central structure includes a shaft sleeve, a central volute around the shaft sleeve, a bearing cavity and a shaft hole supported by the volute envelope structure, and the pressure-bearing surface of the pressure-bearing cover plate is flat. Or the rotating curved surface that coincides with the opening surface of the centripetal guide wheel. The paraxial part has a ring-shaped opening that communicates with the worm. The center worm is a three-dimensional worm that moves radially inward and axially. The beginning is a ring-shaped opening. The end of the tongue on the plane is below the tongue with increased radial and axial coordinates. The cross-sectional area of the worm is proportional to the center angle. The proportionality factor is equal to the volume displacement of the impeller through a unit angle and the average speed of the liquid flow. The ratio is based on the open circular plane. As the cross-sectional area increases linearly, the radial and axial coordinates of the centerline of the bottom of the volute gradually increase, forming a snail bottom slope, which enters the tongue after 360 degrees. Square, followed by the pipe, the cross-sectional shape of the worm also changes with the center angle, starting from the straight segment of the tongue, the first is the semi-axial long semi-ellipse with the long axis shortening on the opening plane, gradually sinking and becoming smooth after becoming a semi-circle Increasing the curvature radius of the lower part, extending along an involute arc sufficient to bypass the curvature change rate of the bearing cavity support structure, into a curved edge quadrangle plus a semicircle shape, until it enters under the tongue, and then deforms while maintaining the cross-sectional area to The circular cross section fits into the pipe.

The centrifugal pump according to claim 6, characterized in that: the impeller is a semi-open circle with a shaft hole (2), a shaft sleeve (3), a blade (5) and a blade groove flow channel (6) Disk-shaped parts, the outside of the sleeve is a ring-shaped suction chamber (4), whose bottom surface may make fluid flow Continued turning curved surface, or flat surface, the latter is equipped with pre-spinner type, the surface of the pre-spinner rim has a turning surface with the same shape. The blade (5) is L-shaped, the front middle part is radial, and the tail is facing It is curved in the tangential direction, and the outer side of the tail is a smooth involute arc cylinder or groove surface. 6 to 12 identical L-shaped blades are evenly distributed on the wheel disc, and a uniformly-distributed lobed channel is formed between the channels. 7) and the middle part (6) are relatively wide, and the cross-sectional area gradually decreases and turns before reaching the exit (8).

 14. The centrifugal pump according to claim 13, characterized in that: a closed impeller is adopted and is closed by adding a front cover (10) on the basis of a half-open high potential ratio impeller, and the front cover has The inner surface of the semi-open type disc is closely matched with the outer surface of the rotating curved surface. Each elbow of the L-shaped blade of the semi-open type foundation structure has 2 to 3 rivet holes or screw holes perpendicular to the disc, and a cover plate. It is riveted with countersunk or flat head rivets (9) or screwed to prevent loosening and fastened to the wheel, or the front cover is connected to the wheel by spot welding.

15. The centrifugal pump according to claim 7, characterized in that: the impeller is provided with an axial inflow pre-spinner, which is composed of two sections of a wheel and elastic sail-type blades of which the number of blades is less than that of the impeller. ) And (15) are sleeved on the impeller suction chamber shaft sleeve, which can rotate independently, and their surfaces kiss each other to form a rotating curved surface to turn the liquid flow. The sail blade (16) forms a curved edge triangle, and its leading edge is suspended on a straight line. On the rigid rib (20), the rib is fixed radially at the impeller blade or the front cover entrance, and two points (18) and (19) on the curved edge of the blade that are close to the bottom surface under the two rims are respectively fixed on the circumference of the bottom surfaces. on prerotator when mounted on the semi-open impeller (31), the rigid ribs ⁽³⁴⁾ at the inlet surface of the radial clamping portion of the blade only Gen.

16. The centrifugal pump according to claim 7, characterized in that the radial inflow pre-spinner comprises a disc-shaped rib support with a sleeve, a sleeve (25), a lower wheel 圏 (21), and an upper wheel 圏 ( 22) Assembled with rigid ribs (23) and elastic sail blades (24), the number of which is less than or equal to the number of impeller blades. The blades are curved triangles, and the leading edge of the blades is suspended on the ribs. The ribs are fixed on the brackets. The shaft sleeve fits statically on the shaft, the blade Points (27) and (29) on the sides of the curve that are close to the bottom surface of the two rims are respectively fixed on the circumferences of the two bottom surfaces, and the two rim slides are sleeved on the bracket shaft sleeve, and the surfaces thereof kiss each other to turn the liquid flow. Rotation surface ₀

 17. The centrifugal pump according to claim 8, characterized in that the high potential ratio impeller is a semi-open type component, or a closed structure using the component as a basic wheel disc, and the wheel edge is circular or sawtooth-shaped. There are L-shaped blades (36) on the plate, and a uniform speed combing blade (37) is set in the wide middle part of the front of the blade groove, forming uniform speed forks (38), (39). The middle part of the front of the comb blade also shows a radial trend, forming a branch entrance. (41) Approaching without reaching the inlet of the blade slot, the tail of the comb blade turns smoothly, and the downstream direction points to the acceleration section of the blade slot, forming a branch outlet. (40) Approaching without reaching the outlet of the blade slot, the area of the outlet from the near pressure surface to near The suction surface decreases.

 18. The centrifugal pump according to claim 17, characterized in that: a pre-spinning average speed high-potential ratio impeller is used, and a constant-speed bifurcation (45) is arranged in the blade groove; the suction chamber is a circular cylindrical shape, and the pre-spinning is installed therein; Device, its two sections of rims (46), (47) are made of self-lubricating material, the sliding sleeve is on the impeller shaft sleeve (50), and its elastic sail blades (48) are hung on rigid ribs such as

(49), the rib is suspended at the entrance of the impeller blade root.

 19. The centrifugal pump according to claim 8 or 17 or 18, characterized in that: the average speed high potential ratio impeller (54), the pre-spinner (55), the mechanical shaft seal (56), and a secondary worm The curved front cover (58), the rear cover (59) and the cantilever shaft form a pre-spinning average speed high-potential ratio secondary worm-type spiral arm pump.

 20. The centrifugal pump according to claim 2 or 9 or 10, characterized in that: at the shaft end of the pump, or a ring cover (68) is used to separate the shaft seal cavity to form a shaft-holding ring-shaped opening (69) and The end cavities are in communication, and the two-phase inflow tube (64) is inserted into the shaft seal cavity. An annular groove is set at the suction end or near the axis of the cavity wall to incorporate a small moving ring that rotates with the impeller.

(74), separating it into a V-shaped ring groove with a distal far end, and a gap (72) on one side communicates with the front cavity (71), and a gap (74) on the other side communicates with the suction chamber, and becomes a V-shaped gas-liquid separation and shunt. Gas blocking gap in the tank, the two-phase inflow pipe (75) communicates with the gap (72), part of the liquid enters the gap (74) from the bottom of the V-shaped groove and returns to the suction chamber (77), and the gas is in the gap (72) The middle float rises to the paraxial axis and flows into the front-end cavity (71) together with the remaining liquid.

 21. The centrifugal pump according to claim 2 or 9, characterized in that: the internal friction reducing driving device of the semi-open impeller cantilever pump is composed of a pressure liquid regulating valve (79), a jet (81), and an ejection gas regulating valve. (80) and connecting thin tube. When the outlet pressure of the pump is higher than the static pressure of the impeller by more than 0.05 MPa ', the pressure liquid is diverted from the outlet. When air is used, the inlet end is connected to the atmosphere and the output two-phase flow is connected to the shaft. Seal the cavity, cool the shaft seal (82), and flow into the rear cavity from the annular cover outlet (83).

 22. The centrifugal pump according to claim 2 or 9, characterized in that the internal friction reducing driving device of the closed impeller centrifugal pump is composed of a pressure liquid regulating valve (92), a jet (90), and an ejection gas regulating valve ( 91), the flow distribution pipes (87) and (88) and the front-end cavity choke gap (93). When the pump outlet pressure is more than 0.05MPa higher than the impeller output static pressure, the pressure liquid is diverted from this outlet, and the output is The phase flow controls the steady-state flow distribution of the front and rear chambers through the flow distribution pipes (88) and (87), where (87) is connected to the shaft seal chamber, and after cooling the shaft seal (85), flows from the annular cover outlet into the rear chamber. (88) is connected to the front end cavity, and is blocked by gas blocking gap (93) to prevent escape. The gas blocking gap (93) is not a V-ring groove but a soft stop at the same position, and is lubricated with two-phase flow. .

 23. The centrifugal pump according to claim 3, characterized in that: the centrifugal guide wheel is composed of a circular cylindrical housing (101), a base plate (103), a shaft sleeve (104), and a guide vane with a decreasing radius of curvature ( 106), the casing is provided with a transfer channel cavity (107) on the shell, and the deceleration and pressure increasing flow path (109) is between the guide vanes, the number of guide vanes is less than the number of impeller blades, and the full deceleration and pressure increasing ratio is equal to the inlet and outlet cutoffs. The area is inversely proportional, and each flow channel converges in the central ring cavity and is constrained by the rotating surface of the sleeve. The liquid flow rotates 90 degrees axially. Or, the central ring cavity is equipped with a radial inflow pre-rotator. .

24. The centrifugal pump according to claim 11, characterized in that the cut-off tongue (118) of the inlet section of the flow guide channel is taken as the starting point, and the center of the interval between the tongues is used as a period to transfer the section to the centrifugal guide wheel. The shape and area of the composite cross section of the flow channel change periodically with the center angle. The change rule is: a. The cross-sectional area increases linearly, and the two parts change in two sections. From the starting point to the positive point of the supercharging section of the previous road (117), the impeller cavity portion linearly increases from the minimum to the maximum, and the guide wheel portion is 0. From this point to the end point, the impeller cavity portion linearly decreases from the maximum value to the minimum value, and the guide wheel portion linearly increases from 0 to the maximum value;

 b. The minimum value of the above composite cross-sectional area is equal to the minimum value of the impeller cavity part, that is, the area of the curved triangle formed by the appearance position of the tongue (118), the curve (119) and the generatrix of the impeller cylindrical surface. The curve (119) consists of two ellipses. The arc is connected to the middle arc. The maximum value of the composite cross-sectional area is equal to its minimum value plus the maximum cross-sectional area of the guide wheel. The latter is equal to the product of the proportional increase of the composite cross-sectional area and the period of the center angle. The maximum cross-sectional area of the cavity is equal to the area of the curved triangle formed by the location of the tongue (118), the curve (120), and the generatrix of the impeller cylindrical surface. The curve (120) is the maximum margin of the cross section of the confluence. It is determined by the coordinates of the end points and the increment of the composite cross-sectional area during the cut-off period of the entrance section;

c. The shape of the two parts is changed in two sections. The impeller cavity part is a curved triangle during the increase in area. The curve edge is a series of intermediate curves starting from the curve (119) to the curve (120). The guide wheel is in the area. During the increase, the shape starts from the initial straight line segment (118) and undergoes the following changes: The long semi-ellipse with (118) as the long axis and the short semi-axis gradually increase, becoming a semi-circle and then moving forward, followed by a rectangle, and moving forward to (II ⁶ ) When the tongue appears at the position, it becomes a pressurized section;

 d. After becoming the pressurized segment, the marginal lines (116) and (118) continue to deform and move forward to form the minimum physical width of the tongue, and the backward side marginal line changes from the tongue forward to the marginal straight line segment to the backward. The curved long semi-ellipses gradually increase in their short semi-axes. After becoming a semicircle, they are translated to the bottom of the runner (117). The front semicircle also moves forward continuously. When it moves to the tangent to the bottom (115), it changes to continuous. The compression becomes a long semi-ellipse, and finally becomes a straight line. During the section's forward deformation, the cross-sectional area expands according to the requirements of deceleration and pressure increase, and the radial coordinate of its centerline also changes.

25. The centrifugal pump according to claim 11, wherein: The guide vane has a starting point of variable curvature that is in contact with the circumferential cavity wall. This starting point is the positive position of the pressurized runner after the transfer. This point determines the radial coordinate of the center of the runner in the transfer section. The section of the runner in the transfer section is divided into impellers. The cavity part and the guide wheel part, the two parts are assembled and joined together. The impeller cavity part of the cross section is wrapped in the impeller by the impeller cover, has a fixed area and shape, and it is responsible for the steering adjustment of the axial surface speed component. It is a cross section of a space between two adjacent tongues that communicates with the impeller cavity. This section independently controls the process of confluence and tangential and axial movement. As the center angle of the flow guide increases, the section of the guide wheel section is The tongue is periodically changed for the starting point and the ending point. The change rule in a cycle is:

 a. The cross-sectional area linearly increases from 0 to the maximum value, and the increase proportionality factor is equal to the design value of the volume of the liquid flow discharged by the impeller through a unit angle divided by the design value of the absolute speed of the liquid flow outlet, or multiplied by a value greater than 1 and less than the guide. The expansion coefficient of the minimum expansion ratio of the wheel pressurized runner. The maximum cross-sectional area is equal to the increase of the proportional coefficient multiplied by the center angle of the inlet section of the runner. When the increased proportional coefficient includes the expansion coefficient factor, the flow section of the transfer section has a deceleration increase. Pressure function

 b. The cross section starts from the starting straight line segment (128) and undergoes various shape changes. First, the long semi-ellipse with (128) as the long axis, and the short semi-axis gradually increases: after becoming a semicircle, it becomes a semicircle with continuous margins. Move forward to form a rectangular cross section followed by a front semicircle; when the tongue appears, the semicircle moves to the (126) position, and the flow passage of the transfer section is separated from the impeller cavity to become a pressurized flow passage;

c. After the tongue appears and isolates the impeller cavity, the flow channel defined by (126) and (128) becomes a pressurized flow channel and continues to move forward and deform. After a small amount of forward movement leaves the minimum physical width of the tongue, The backward side marginal line changes from a straight line segment to a long curved semi-ellipse that is curved backwards. The short semi-axis is continuously lengthened to become a semicircle and then changed to translation until the last point enters the bottom surface of the runner (127). During this process, the The forward marginal semicircle continues to move forward until it is tangent to the bottom plane 125 of the guide wheel. Instead, it continuously compresses the semicircle into a semi-ellipse, and finally becomes a straight line penetrating the bottom plane. The contour line or center of mass should move at a high speed. An appropriate percentage of the corresponding movement speed during the confluence, for example, greater than 50%, so that the cross-sectional area and strength of the tongue can be continuously increased. During the forward-flow section of the booster channel, its area changes according to the deceleration and boosting requirements.

 26. The centrifugal pump according to claim 12, and the technical expansion of the centrifugal pump modular combination method according to claim 4 is used for technical expansion, characterized in that: the symmetrical end cover module is composed of a pressure bearing cover with an assembly stop (141) Plate (142), three-dimensional wormway (146) on the cover plate and its annular entrance and exit (143), a straight section pipe that is in contact with the wormway (146), a sleeve (149) supported by the wormway structure, and a bearing cavity (145) and other structural components, the worm trajectory has a constraint feature that the cross-sectional area is proportional to the center angle, and the friction surface and acceleration are optimized. The purpose is to expand and standardize its functional design and usage, including: a. Utilizing its ring shape The interface and the internal characteristics of the three-dimensional worm are compatible and constrained by the three-dimensional motion. Construct or adaptively generate the connection boundary conditions such as the shunt, confluence, rotation, and axial surface steering of the multi-flow path of the impeller and the guide wheel. Meet the requirements of centrifugal guide wheel outlet connection, make it universal for single-stage pumps and multi-stage pumps, and be used as a universal flow field boundary module for single-stage and multi-stage pumps. To support the conservative loop connection mode described;

 b. Utilize the characteristics of its circular interface and internal compatibility of three-dimensional worm trajectory to constrain three-dimensional motion, and expand the compatibility of technology design compatibility of inflow and outflow directions with reciprocal and divergent confluence properties, and use it as a conservative loop connection mode. Front-to-back flow field symmetrical boundary module;

 c. Limit and optimize the spatial and temporal rate of change of the flow velocity by using the circular interface and the directional compatibility within the three-dimensional volute and the continuity of the three-dimensional motion, so as to minimize it and use it as a stable and low-loss feature. Flow field boundary module to achieve high performance in conservative loop connection mode;

d. Utilizing the rotational symmetry of its ring-shaped interface and assembly stop, and the relative positioning of the worm and the lead-out tube based on the relative angle of the separating tongue, based on the functions of independently changing the angle of the front and rear cover outlet tubes to support modularity Assembly required assembly structure and function Extended design

 e. Utilizing its integrated structural features such as pressure plates and bearing housings, it is determined as a moldable single-part functional component in terms of technology and process design. After the above technology is expanded, the structure that supports the connection mode is constructed. The end cap module with standardized assembly dimensions and interface parameters expands the value range of true small size, simple design, low cost, and powerful functions.

 27. The centrifugal pump according to claim 3 or 11 or 23 or 24 or 25, which adopts the modular combination method according to claim 4, characterized in that: it comprises a centripetal booster module, which is guided by the centripetal guide The impeller, impeller and impeller cavity cover are axially combined, or are equipped with other functional accessories, with standardized interface parameters and assembly dimensions. The centrifugal guide wheel is integrally manufactured by molding, and its cavity side plane or rotating curved surface Cooperate with the impeller to form a gap. The outer edge of the cavity side has a forward marginal curved surface of the transfer section flow channel designed according to the parameters of the impeller. The stepped shell has an assembly stop. The middle partition plate connected to the shell serves as the guide vane support substrate. It is used to isolate the impeller cavity and the guide wheel cavity and bear the pressure difference between them. The impeller cavity cover plate is a molded weight-reducing structural part. The cavity-side rotating curved surface forms a clearance fit with the impeller, and the outer edge of the cavity side is designed according to the impeller parameters. The segmental flow path has a backward marginal surface, so that the transfer segment flow path is fully constrained and connected in series with the pressurized flow path. During assembly, the guide wheel, impeller, and impeller cavity cover plate are sequentially installed. The three are respectively positioned by the casing stop, the shaft and the impeller cavity of the guide wheel. During operation, the liquid flow flows axially from the module inlet into the rotating impeller flow channel, and the normal force work of the blade is accelerated along the way and the centrifugal force work is integrated. Increase the specific energy, flow into the guide wheel through the flow channel of the transfer section, decelerate and increase the pressure, and then rotate 90 degrees from the paraxial annular port to flow out of the module.

 28. The centrifugal pump according to claim 27, comprising: a semi-open impeller centripetal booster module, the module is composed of a centripetal guide wheel (157), a semi-open impeller (154), and an impeller cavity cover plate (152) An axial combination.

29. The centrifugal pump according to claim 27, comprising: a centrifugal booster module with a closed impeller, the module comprising a centrifugal guide wheel (167), a closed impeller (164) and The impeller cavity cover plate (162) is axially combined, and the space occupied by the impeller cover is cut out from the impeller cavity cover (162) so that the impeller flow passage and the outlet flow passage are in abutment with each other.

 30. The centrifugal pump according to claim 27, comprising: a friction-reducing closed-type impeller centripetal booster module, the module comprising a centrifugal guide wheel (179), a closed-type impeller (180), and an impeller cavity cover ( 175) and V-shaped groove air gap ring cover (171), V-shaped groove moving ring (172), inflatable drive two-phase inflow pipe (174), front and rear cavity pressure equalization holes (177) in the fixed rivet of the front cover The internal friction reducing parts or structure are combined, wherein the space occupied by the impeller cover is cut out from the impeller cavity cover (175) so that the flow path of the impeller is in contact with the outer flow channel of the outlet, and the pressure equalizing hole (177) makes The front and rear cavities communicate with each other, and the annular groove on the cover plate (175) of the impeller cavity, the annular cover plate (171) and the rotating movable ring (172) form a V-shaped air gap.

 31. The centrifugal pump according to claim 27, comprising a centrifugal booster module with a half-open average speed high potential ratio impeller, the module is composed of a centripetal guide wheel (188), a half-open average speed high potential The specific impeller (183) and the impeller cavity cover (182) are axially combined.

 32. The centrifugal pump according to claim 27, comprising: a closed-type average speed high-potential ratio impeller centripetal booster module, which is composed of a centripetal guide wheel (198), a semi-open type average speed high-potential ratio The impeller (193) and the impeller cavity cover plate (192) are axially combined. The impeller cavity cover (192) is cut out of the space occupied by the impeller cover, and the blade groove flow channel is in contact with the outlet outer flow channel.

 33. The centrifugal pump according to claim 27, comprising: a pre-spinning closed-speed high-potential ratio impeller centrifugal booster module, which is closed by a centripetal guide wheel (208) and an average-speed high-potential ratio block. The impeller (204), the pre-rotator (201), and the impeller cavity cover plate (202) are axially combined. The impeller cavity cover (192) has a space occupied by the impeller cover.

34. The centrifugal pump according to claim 27, comprising: a centrifugal booster module for reducing friction, a closed-type, high-velocity, high-potential ratio impeller, which is composed of a centrifugal guide wheel (219), a closed-type, high-velocity, high-potential Than impeller (220), impeller cavity cover (215) and V-groove The gap annular cover plate (211), the V-shaped groove moving ring (212), the two-phase inflow pipe (214), and the front and rear cavity pressure equalizing holes (217) in the front cover rivet are combined, and the impeller cavity cover (215 The space occupied by the impeller cover is dug up.

 35. The centrifugal pump according to claim 27, comprising: a pre-spinning friction-reducing closed-type average speed high-potential ratio impeller centrifugal booster module, the module is composed of a centripetal guide wheel (228), a closed-type average speed High potential ratio impeller (230), impeller cavity cover plate (225), axial inflow pre-rotator (221), V-groove dynamic ring gas blocking structure (222), two-phase inflow pipe (223), front and rear cavity Composed of pressure equalizing holes (227).

 36. The centrifugal pump according to claim 27, comprising: an ultra-reduced anti-friction, pre-spinned, closed-velocity high-potential ratio impeller centrifugal booster module, the module comprising a centrifugal guide wheel

(240), a closed average speed high potential ratio impeller (236) with an impeller cover plate (234) extending and including a cross section of the impeller cavity section of the transfer channel, an impeller cavity cover plate (235), and an axial incoming flow pre-spin (231) and ultra-reduced friction assembly (232), two-phase inflow pipe (233) assembly.

 37. The centrifugal pump according to claim 12 or 26, which adopts the modular combination method according to claim 4, characterized in that: it comprises two variable-angle outlet pipe symmetrical end cover modules and a centripetal booster module, The two modules are fully interchangeable according to the corresponding child specifications, or the parameters of the medium, the maximum speed, the maximum temperature, and the maximum pressure resistance are verified by checking the table according to the corresponding parent specifications. "Circular port with ring-shaped inflow and outflow" connection mode combines three modules axially, that is, a symmetrical cover with variable angle and a variable-angle outlet tube centrifugal single-stage centrifugal pump with modular interchangeability. The combination refers to the connection and coordination in the design. , Assembly in production and repair in use, interchangeability covers these processes.

38. The centrifugal pump according to claim 28 and 37, comprising: a semi-open impeller centripetal booster module and two variable angle outlet tube symmetrical end cover modules, the former being a semi-open impeller (245 ), The impeller cavity cover (243) and the centrifugal guide wheel (247)), the latter is used as the front cover (241) and the rear cover (250), respectively, axially combined to have a mold Block interchangeable symmetrical cover variable angle outlet tube semi-open impeller centrifugal booster single-stage centrifugal pump.

 39. The centrifugal pump according to claim 29 and 37, comprising: a closed-type impeller centripetal booster module and two variable-angle outlet pipe symmetrical end cover modules, the former by a closed-type impeller (255), The impeller cavity cover (253) and the centripetal guide wheel (257) are used as the front cover (251) and the rear cover (260), respectively, and are axially combined to form a symmetrical cover with modular interchangeability and variable angle outlet pipe closure. Centrifugal pump with centrifugal impeller.

 40. The centrifugal pump according to claim 30 and 37, comprising: a friction-reducing closed-type impeller centripetal booster module and two variable-angle outlet pipe symmetrical end cover modules, the former being a closed-type impeller (268 ), The impeller cavity cover (265), the centripetal guide wheel (267) and the air gap (263), the two-phase inflow pipe (264), the front and rear cavity communication pressure equalization holes (266), the latter is used as the front The cover (261) and the back cover (270) are axially combined to form a symmetrical cover variable angle outlet tube anti-friction closed impeller centrifugal single-stage centrifugal pump with module interchangeability.

41. The centrifugal pump according to claim 31 and 37, comprising: a semi-open type constant velocity high potential ratio impeller centripetal booster module and ^two variable angle outlet tube symmetrical end cover modules, the former consists of a half Open type average speed high potential ratio impeller (274), impeller cavity cover (273), centripetal guide wheel (277), the latter is used as the front cover (271) and the rear cover 80), which are axially combined to have a module Interchangeable symmetrical cover variable angle outlet pipe semi-open type uniform speed high potential ratio centrifugal single-stage centrifugal pump with centrifugal booster.

42. The centrifugal pump according to claim 32 and 37, comprising: a closed-type high-potential ratio impeller centripetal booster module and two variable-angle outlet symmetrical end-cap modules, the former being a closed type The uniform speed high potential ratio impeller (284), the impeller cavity cover ( ²⁸³ ), and the centripetal guide wheel (287), the latter is used as the front cover (281) and the rear cover (290), which are axially combined to have module mutual A single-stage centrifugal pump with centrifugal pump with centrifugal pressure and a centrifugal centrifugal pump.

43. The centrifugal pump according to claim 33 and 37, comprising: a pre-spin closed closed-velocity high-potential ratio impeller centrifugal booster module and two variable-angle outlet pipes symmetrical End cover module, the former consists of a closed-type average speed high potential ratio impeller (295), a pre-rotator (293) installed in the impeller suction chamber, an impeller cavity cover (294), a centripetal guide wheel (297), and the latter They are respectively used as the front cover (291) and the rear cover (300), and are axially combined to form a symmetrical cover variable angle outlet pre-spinning closed-type high-potential ratio impeller centrifugal single-stage centrifugal pump with modular interchangeability.

 44. The centrifugal pump according to claim 34 and 37, comprising: a friction-reducing closed-type constant-velocity high-potential-ratio impeller centrifugal booster module and two variable-angle outlet-tube symmetrical end-cap modules, the former by Closed average speed high potential ratio impeller (308), impeller cavity cover (305), centripetal guide wheel (307) and choke gap (303), two-phase inflow pipe (304), front and rear cavity pressure equalization holes (306 ), The latter is used as the front cover (301) and the rear cover (310), axially combined to form a symmetrical cover with modular interchangeability, variable angle outlet tube, reduced friction, closed average speed, high potential ratio than the impeller Single-stage centrifugal pump.

 45. The centrifugal pump according to claim 35 and 37, comprising: a friction-reducing pre-spinning type average speed high potential ratio impeller centrifugal booster module and two variable angle outlet pipe symmetrical end cover modules, The former consists of impeller (318) with closed average speed and high potential ratio, and impeller cavity cover.

(315), a centrifugal guide wheel (317), a pre-rotator (312) and a choke gap (313), a two-phase inflow pipe (314), and a pressure equalizing hole (316) in the front and rear chambers, the latter being used as The front cover (311) and the rear cover (320) are axially combined to form a symmetrical cover with variable module angles, variable angle outlet pipes, anti-friction, pre-rotating, closed-type, high-potential, single-stage centrifugal pump with centrifugal pressure.

 46. The centrifugal pump according to claim 36 and 37, comprising: a super-reduction anti-friction pre-rotational closed-speed high-potential ratio impeller centrifugal booster module and two variable-angle outlet pipe symmetrical end cover modules The former consists of an impeller cover (326) with a closed average speed high potential ratio impeller (327), an impeller cavity cover (325), a centripetal guide wheel (329), and a pre-rotator ( 322) and air-barrier gap (323), front-end cavity two-phase inflow tube (324), rear-end cavity two-phase inflow tube (328), the latter being used as the front cover (321) and the rear cover, respectively

(330), axially combined into a symmetric cover variable angle outlet pipe with super-friction, pre-rotational closed-type, high-potential ratio impeller centrifugal single-stage centrifugal pump with module interchangeability.

47. The centrifugal pump according to claim 12 or 26, which adopts the modular combination method described in claim 4, characterized in that it comprises: 2 variable angle outlet pipe symmetrical end cover modules and a plurality of 64 multi-directional Cardiac boost module, the two modules are fully interchangeable according to the corresponding child specifications, or the parameters of the medium, the maximum speed, the maximum temperature, and the maximum pressure resistance are verified by a look-up table according to the corresponding parent specifications, of which the highest The interchangeability of pressure resistance or axial segmentation is established. According to the connection mode of "liquid flow flowing in and out from the paraxial annular mouth with ring volume", the symmetrical end cap module is divided into front and rear caps, which will pressurize centripetally. The modules are axially connected in series in sequence, and all the modules are axially combined, that is, a symmetrical cover with variable angles and a centrifugal multi-stage centrifugal booster pump with modular interchangeability are formed. Interchangeability, interchangeability covers these processes.

 48. The centrifugal pump according to claim 28 and 47, comprising: a plurality of semi-open impeller centrifugal booster modules and two variable angle outlet tube symmetrical end cover modules, the former being a semi-open impeller (334 ), The impeller cavity cover (333) and the centripetal guide wheel (335), the latter is used as the front cover (332) and the rear cover (340), respectively, axially combined to form a symmetrical cover with modular interchangeability. Tube semi-open impeller centripetal booster multistage centrifugal pump.

 49. The centrifugal pump according to claim 29 and 47, comprising: a plurality of closed impeller centripetal pressure boosting modules and two variable angle outlet tube symmetrical end cover modules, the former being a closed impeller (344), The impeller cavity cover (343) and the centripetal guide wheel (345) are used, the latter is used as the front cover (342) and the rear cover (349), which are axially combined to form a symmetrical cover with modular interchangeability and variable angle outlet pipe closure. Centrifugal multi-stage centrifugal pump with centrifugal impeller.

50. The centrifugal pump according to claim 30 and 47, comprising: a plurality of anti-friction closed-type impeller centrifugal booster modules and two variable-angle outlet pipe symmetrical end cover modules, the former being a closed impeller (353 ), The impeller cavity cover (352), the centripetal guide wheel (354), the air gap (355), the two-phase inflow pipe (356), the front and rear cavity communication pressure equalization holes (358), the latter is used as the front The cover (342) and the back cover (349) are axially combined to form a symmetrical cover with variable module interchangeability, a variable angle outlet tube, a friction reducing closed impeller, and a centrifugal multi-stage centrifugal pressurization. Pump.

 51. The centrifugal pump according to claim 31 and 47, comprising: a plurality of semi-open type constant speed high potential ratio impeller centrifugal booster modules and two variable angle outlet tube symmetrical end cover modules, the former consists of half Open type average speed high potential ratio impeller (364), impeller cavity cover (363) and centripetal guide wheel (365), the latter is used as the front cover (362) and the rear cover (369), which are axially combined to have Module interchangeable symmetrical cover variable angle outlet pipe semi-open type uniform speed high-potential ratio multi-stage centrifugal pump with centrifugal booster.

 52. The centrifugal pump according to claim 32 and 47, comprising: a plurality of closed-type high-potential ratio impeller centrifugal booster modules and two variable-angle outlet pipe symmetrical end cover modules, the former being a closed type The average speed high potential ratio impeller (374), the impeller cavity cover (373) and the centripetal guide wheel (375), the latter is used as the front cover (372) and the rear cover (379), which are axially combined to have module mutual Multi-stage centrifugal pump with centrifugal pump and centrifugal pump with centrifugal pump, closed-end, closed-velocity, high-potential than the impeller.

 53. The centrifugal pump according to claim 33 and 47, comprising: a plurality of pre-rotated closed-type high-potential ratio impeller centrifugal booster modules and two variable-angle outlet pipe symmetrical end cover modules, the former by Closed average speed high potential ratio impeller (384), pre-rotator (386), impeller cavity cover (383) and centripetal guide wheel (385), the latter is used as the front cover (382) and the rear cover (389) ), Axially combined into a modular cover interchangeable symmetrical cover variable angle outlet pipe pre-spinning closed-velocity high-potential ratio impeller centrifugal booster multi-stage centrifugal pump.

 54. The centrifugal pump according to claim 34 and 47, comprising: a plurality of friction-reducing closed-type constant-velocity high-potential-ratio impeller centrifugal booster modules and two variable-angle outlet pipe symmetrical end cover modules, the former by Closed average speed high potential ratio impeller (394), impeller cavity cover (393), centripetal guide wheel (395) and choke gap (396), two-phase inflow pipe (397), front and rear cavity communication pressure equalization holes ( 398), the latter is used as the front cover (392) and the rear cover (399) respectively, axially combined into a symmetrical cover with modular interchangeability, variable angle outlet tube, reduced friction, closed type, and high average potential increase than the impeller. Pressure multistage centrifugal pump.

55. The centrifugal pump according to claim 35 and 47, wherein: Anti-friction pre-spinning closed-type high-speed ratio impeller centripetal booster module and 2 variable-angle outlet pipe symmetrical end cover modules, the former consists of a closed-type high-speed ratio impeller (406) and prerotator (404) , Impeller cavity cover (403), centripetal guide wheel (405) and air gap (408), two-phase inflow tube (407), front and rear cavity communication pressure equalization holes (409), the latter is used as the front cover, respectively (402) and rear cover (410) are axially combined to form a symmetrical cover variable angle outlet tube with anti-friction and pre-spin closed high-potential ratio centrifugal multi-stage centrifugal pump.

 56. The centrifugal pump according to claim 36 and 47, comprising: a plurality of super-reduction anti-friction pre-spinning type average speed high potential ratio impeller centrifugal booster modules and a variable-angle outlet pipe symmetrical end cover module The former consists of a closed average speed high potential ratio impeller (414), an impeller (411), an impeller cavity cover (413), and a centripetal guide wheel (414 415) and choke gap (416), front-end cavity friction reduction driving two-phase inflow tube

(417). The rear end cavity is a friction-reducing driving shield insertion tube (419), which is used as a front cover (412) and a rear cover (420), respectively. The axial combination forms a symmetrical cover with modular interchangeability. Multi-stage centrifugal pump with super-friction tube pre-spinning closed-type average speed high potential ratio centrifugal booster with centrifugal booster.

 57. The centrifugal pump according to claim 47, comprising: a plurality of pre-spinning double-half-type average speed high-potential ratio impeller centrifugal booster modules and two variable-angle outlet pipe symmetrical end cover modules, The former consists of a semi-open type, high-potential ratio impeller (424), and an impeller cavity cover.

(423), a semi-open centrifugal guide wheel (425) and a pre-spinner (428), the latter being used as a front cover (422) and a rear cover (429), respectively, and axially combined into a module interchangeable Symmetrical cover variable angle outlet pipe double semi-open type uniform speed high potential ratio centrifugal booster pump with centrifugal booster.