Turbo compressor system and operation method for turbo compressor system

Abstract

To provide a turbo compressor system and an operation method for the turbo compressor system that can realize capacity control having high reliability, high efficiency, and a wide operation range.SOLUTION: A turbo compressor system has: a turbo compressor having a driving machine, a rotating shaft for transmitting the rotation of the driving machine, and impellers attached to the rotating shaft and for compressing gas; and a suction pipe for supplying the gas into the turbo compressor, and branching into a first suction pipe and a second suction pipe upstream of the turbo compressor. Suction flow passages for introducing the gas into a first-stage impeller of the turbo compressor have suction flow passages having a first suction flow passage connected to the first suction pipe and for introducing the gas into the turbo compressor without swirling it, and a second suction flow passage connected to the second suction pipe and for introducing the gas into the turbo compressor after swirling it. The turbo compressor system comprises at least one first flow adjustment mechanism capable of continuously adjusting a flow ratio of the gas passing through the first suction flow passage and the second suction flow passage of the turbo compressor.SELECTED DRAWING: Figure 1

Description

[Technical Field]

[0001] The present invention relates to a turbocompressor system and a method of operating a turbocompressor system. [Background technology]

[0002] Various types of turbo compressors are used in various plants to boost the pressure of process gases. A turbo compressor is configured to draw in process gas through a suction passage, compress and boost the process gas using an impeller attached to a rotating shaft, and then discharge the gas through a discharge passage. Centrifugal compressors are often required to operate stably over a wide range, from the rated point to low airflow.

[0003] An example of a capacity control method for a turbo compressor is described in Japanese Patent Application Laid-Open Publication No. 2010-236401 (Patent Document 1). The centrifugal compressor described in this patent document is a radial suction centrifugal compressor having an inlet passage through which process gas flows in from a radial direction, which is approximately perpendicular to the rotation axis. In the inlet passage upstream of the first-stage impeller, multiple movable inlet guide vanes with a drive mechanism and variable mounting angle are installed. By changing the mounting angle of these movable inlet guide vanes, the angle of swirl (pre-swirl) imparted to the first-stage impeller of the centrifugal compressor in the same direction as the impeller rotation direction is gradually increased, thereby lowering the head of the first-stage impeller and shifting the stall flow rate of the first-stage impeller to the low flow rate side, thereby reducing the suction gas volumetric flow rate of the centrifugal compressor. In this case, the pre-swirl allows the compressor head to be efficiently lowered while suppressing an increase in pressure loss, thereby significantly reducing the shaft power on the low flow rate side. [Prior art documents] [Patent documents]

[0004] [Patent Document 1] Japanese Patent Application Laid-Open No. 2010-236401 Summary of the Invention [Problem to be solved by the invention]

[0005] In the centrifugal compressor described in Patent Document 1, the flow rate of the centrifugal compressor is controlled by driving a movable inlet guide vane installed upstream of the first-stage impeller to impart pre-rotation to the first-stage impeller. This allows for more efficient capacity control with a wider stable operating range at low airflow rates than capacity control methods that use compressor suction throttling. However, this requires the compressor to be equipped with a drive mechanism for changing the mounting angle of the movable inlet guide vane, which increases the compressor's structural complexity and size. Furthermore, depending on the specifications of the plant in which the turbo compressor is installed, the compressor's flow rate may need to be controlled frequently. In such cases, the inlet guide vanes must be driven repeatedly at high frequencies, which reduces reliability.

[0006] An object of the present invention is to provide a turbo-compressor system and an operating method for the turbo-compressor system that are highly reliable and capable of realizing highly efficient capacity control over a wide operating range. [Means for solving the problem]

[0007] A turbo-compressor system of the present invention is characterized in that it includes: a turbo-compressor having a driver, a rotating shaft that transmits rotation of the driver, and an impeller that is attached to the rotating shaft and compresses gas; an suction pipe that supplies gas to the turbo-compressor, the suction pipe branching into a first suction pipe and a second suction pipe upstream of the turbo-compressor; an suction passage that introduces gas to a first-stage impeller of the turbo-compressor, the suction passage having: a first suction passage connected to the first suction pipe that introduces the gas into the turbo-compressor without imparting swirl to the gas; and a second suction passage connected to the second suction pipe that introduces the gas into the turbo-compressor with imparting swirl to the gas; and at least one first flow rate adjustment mechanism that can continuously adjust a flow rate ratio of gas passing through the first suction passage and the second suction passage of the turbo-compressor.

[0008] Alternatively, in a method of operating a turbo compressor system of the present invention, the turbo compressor system includes a fixed speed driver, and when reducing an suction gas flow rate from a rated gas flow rate, the method includes fully closing a valve aperture of the second flow control mechanism, starting from an operating state corresponding to a rated operation state in which a flow rate ratio between the first suction passage and the second suction passage is at a maximum and the gas flow rate passing through the first suction passage is greater, adjusting the valve aperture of the first flow control mechanism to gradually reduce the gas flow rate of the first suction passage while increasing the gas flow rate of the second suction passage, and then fully closing the first suction passage, fully opening the second suction passage, and applying pre-swirl to reduce the suction gas flow rate passing through the suction piping to a minimum stable operating point that can be achieved by the turbo compressor.

[0009] Alternatively, in a method of operating a turbo compressor system of the present invention, the turbo compressor system includes a driver that is a variable speed driver, and when reducing the suction gas flow rate from a rated gas flow rate, the method includes: fully closing a valve aperture of the second flow control mechanism, starting from an operating state that corresponds to a rated operation state in which the flow rate ratio between the first suction passage and the second suction passage is at a maximum and the gas flow rate passing through the first suction passage is greater; adjusting the valve aperture of the first flow control mechanism to gradually reduce the gas flow rate of the first suction passage while increasing the gas flow rate of the second suction passage; nearly fully closing the first suction passage; fully opening the second suction passage to further reduce the suction gas flow rate by applying pre-swirl; and then reducing the rotation speed of the driver to reduce the suction gas flow rate passing through the suction piping to a minimum stable operating point that can be achieved by the turbo compressor. [Effects of the Invention]

[0010] According to the present invention, it is possible to provide a turbo compressor system and an operating method for a turbo compressor system that are highly reliable and capable of realizing highly efficient capacity control over a wide operating range. [Brief explanation of the drawings]

[0011] [Figure 1] 1 shows a turbo compressor system according to a first embodiment. [Figure 2]FIG. 2 is a diagram illustrating the flow rate-discharge pressure characteristic curve of a turbo compressor and the plant operating point in the case of a fixed speed drive machine. [Figure 3] FIG. 2 is a diagram illustrating a flow rate-discharge pressure characteristic curve and a plant operating point of a turbo compressor in the case of a variable speed drive machine. [Figure 4] FIG. 3 is a diagram showing a modified example of the turbo compressor system of the first embodiment. [Figure 5] FIG. 10 is a diagram illustrating a turbo compressor system according to a second embodiment. [Figure 6] FIG. 10 is a longitudinal sectional view of a single-shaft multi-stage centrifugal compressor according to a third embodiment. [Figure 7] FIG. 7 is a view taken along line II in FIG. 6. [Figure 8] FIG. 7 is a cross-sectional view taken along line II-II in FIG. 6. [Figure 9] FIG. 7 is a cross-sectional view taken along line III-III in FIG. 6. [Figure 10] FIG. 10 is a diagram showing a modified example of the turbo compressor system according to the third embodiment. [Figure 11] FIG. 10 is a diagram of a turbo compressor system according to a fourth embodiment. [Figure 12] FIG. 10 is a diagram showing another example of a turbo compressor system according to the fourth embodiment. [Figure 13] FIG. 10 is a diagram showing another example of a turbo compressor system according to the fourth embodiment. [Figure 14] FIG. 10 is a diagram showing another example of a turbo compressor system according to the fourth embodiment. DETAILED DESCRIPTION OF THE INVENTION

[0012] Specific embodiments of a turbo compressor system according to the present invention will be described below with reference to the accompanying drawings. In the various drawings, identical reference numerals denote identical or corresponding components. In some of the embodiments described below, a single-shaft multi-stage centrifugal compressor will be used as the turbo compressor, but the present invention can be similarly applied to systems having other types of turbo compressors. [Example]

[0013] A specific embodiment 1 of the turbo compressor system will be described with reference to Figs. 1 to 4. Fig. 1 is a diagram of the turbo compressor system in this embodiment 1. Fig. 2 is a diagram illustrating a flow rate-discharge pressure characteristic curve of the turbo compressor and a plant operating point when a turbo compressor system using a fixed speed driver is applied in this embodiment 1. Fig. 3 is a diagram illustrating a flow rate-discharge pressure characteristic curve of the turbo compressor and a plant operating point when a turbo compressor system using a variable speed driver is applied in this embodiment 1. Fig. 4 is a diagram illustrating a modified example of the turbo compressor system in this embodiment 1.

[0014] As shown in FIG. 1, the turbo compressor system 1000 in this embodiment includes a turbo compressor 100 having a driver 21, a rotating shaft 2 that transmits the rotation of the driver 21, and an impeller 1 that is attached to the rotating shaft 2 and compresses gas 25.

[0015] A suction pipe 22 is installed upstream of the turbo compressor 100 to supply gas 25 to the turbo compressor 100. The suction pipe 22 is configured to branch into a first suction pipe 22a and a second suction pipe 22b from a branching portion 22c upstream of the turbo compressor 100.

[0016] Downstream of the first suction pipe 22a and the second suction pipe 22b, suction passages are installed that introduce gas 25 into the first stage impeller 1b installed inside the turbo compressor 100, with the suction passage connected to the first suction pipe 22a being the first suction passage 7 and the suction passage connected to the second suction pipe 22b being the second suction passage 8. In this embodiment, the first suction passage 7 has some kind of passage shape that allows the gas 25 to be introduced into the turbo compressor 100 without imparting pre-swirl to the gas 25. On the other hand, the second suction passage 8 has some kind of passage shape that allows the gas 25 to be introduced into the turbo compressor 100 while imparting pre-swirl to the gas 25.

[0017] A first flow rate adjustment mechanism 23 is installed in the first suction pipe 22a and the second suction pipe 22b to continuously adjust the flow rate ratio of the gas 25 passing through the first suction passage 7 and the second suction passage 8. The first flow rate adjustment mechanism 23 is configured as a flow rate adjustment valve 23a installed in the first suction pipe 22a and a flow rate adjustment valve 23b installed in the second suction pipe 22b. It is equipped with a flow rate control device 40 that controls the flow rate adjustment valve 23a and the flow rate adjustment valve 23b. However, other types of flow rate adjustment mechanisms can also be used. In addition, a discharge pipe 24 is installed to discharge the fluid compressed by the turbo compressor 100.

[0018] Next, the operation of the turbo compressor system of this embodiment will be described. First, FIG. 2 shows the flow rate-discharge pressure characteristic curve of the turbo compressor and the plant operating point when the turbo compressor system 1000 of this embodiment is applied, in which the driver 21 is fixed speed. In the case of a fixed-speed driver, the capacity control means of the turbo compressor 100 is valve opening adjustment of the flow rate control valves 23a and 23b, which are the first flow rate adjustment mechanism 23. Note that this figure assumes a case in which the discharge pressure Pd of the turbo compressor 100 required by the plant specifications is a constant Pd,spec regardless of the flow rate of the gas 25 flowing within the plant (plant flow rate Q), but plants with other system resistance characteristics may also be used.

[0019] First, consider the case where the turbo compressor system 1000 is operated so that the plant flow rate Q is equal to the rated flow rate Qspec. Figure 2 shows a curve 29a, which is the flow rate-discharge pressure characteristic curve of the turbo compressor 100 in this case, and a plant operating point 31, which is the intersection of the curve 29a, which is the flow rate-discharge pressure characteristic curve, and the specified discharge pressure Pd,spec. In this operating state, all of the gas 25 passes only through the first suction passage 7 so that the gas 25 is drawn into the turbo compressor 100 without pre-swirl. Therefore, when the first flow control mechanism 23 is configured as two flow control valves 23a and 23b as shown in Figure 1, the flow control valve 23a is fully open and the flow control valve 23b is fully closed.

[0020] When the plant flow rate is gradually reduced from the rated flow rate Qspec, the openings of the two flow control valves 23a and 23b are adjusted to gradually reduce the flow rate of the gas 25 passing through the first suction passage 7, while gradually increasing the flow rate of the gas 25 passing through the second suction passage 8 at a rate that is smaller than the decrease in the flow rate of the gas 25 passing through the first suction passage 7. As a result, the plant flow rate Q is gradually reduced while the pre-swirl angle imparted to the gas 25 sucked into the turbo compressor 100 is increased, thereby gradually shifting the plant operating point 31 to the lower flow rate side. An example of the flow rate-discharge pressure characteristic curve of the turbo compressor at this time is shown by curve 29b in Figure 2.

[0021] Next, consider a case where the turbo-compressor system 1000 is operated at a flow rate point Qc where the plant flow rate is equal to the minimum stable operating flow rate that the turbo-compressor 100 can achieve. When the turbo-compressor 100 is operated at the minimum stable operating flow rate point, all of the gas 25 drawn into the turbo-compressor 100 passes through the second suction passage 8, which can impart pre-swirl, and the pre-swirl angle imparted to the gas 25 is at its maximum.

[0022] 1, the first flow rate adjustment mechanism 23 is configured as two flow rate adjustment valves 23a and 23b, and the flow rate adjustment valve 23a is fully closed and the flow rate adjustment valve 23b is fully open. When the turbo compressor system 1000 is in this operating state, the turbo compressor 100 operates at the minimum stable operating flow rate that can be achieved by the equipment alone, and the plant operating point 31 coincides with the intersection of curve 29c, which is the flow rate-discharge pressure characteristic curve of the turbo compressor 100, and the specified discharge pressure Pd,spec.

[0023] Note that a line connecting the minimum stable operating flow rate points of the turbo compressor 100 under various conditions, such as when there is no pre-swirl as shown by curve 29a, when an arbitrary pre-swirl angle is applied as shown by curve 29b, and when the pre-swirl angle is maximum as shown by curve 29c, is generally called a surge line 30. The surge line in Figure 2 is the surge line 30 when the pre-swirl angle changes, and is shown as surge line 30a to distinguish it from the surge line when the rotation speed changes.

[0024] Next, Figure 3 shows the flow rate-discharge pressure characteristic curve of the turbo compressor and the plant operating point when a turbo compressor system 1000 is applied in which the driver 21 of this embodiment is variable speed. Using a variable speed driver requires additional equipment such as an inverter, which increases costs. On the other hand, the rotation speed can be gradually increased when starting the turbo compressor 100, reducing starting torque and improving equipment reliability. The rotation speed can be adjusted arbitrarily, allowing for more precise capacity control than when adjusting the valve opening alone.

[0025] In the case of a variable speed drive machine, the capacity control means of the turbo compressor 100 is the valve opening adjustment of the flow rate adjustment valves 23a and 23b, which are the first flow rate adjustment mechanism 23, and the rotation speed of the drive machine 21.

[0026] In Fig. 3, in the rated operating state where Q = Qspec, the compressor rotation speed is the rated rotation speed, and all of the gas 25 passes only through the first suction passage 7, i.e., the flow rate control valve 23a shown in Fig. 1 is fully open, and the flow rate control valve 23b is fully closed. In this case, the flow rate-discharge pressure characteristic curve of the turbo compressor 100 is shown by curve 29a, and the operating point of the plant is shown by plant operating point 31, both in Fig. 3.

[0027] When a variable-speed drive is used, two types of capacity control methods are available for gradually reducing the plant flow rate from the rated flow rate Qspec: adjusting the valve opening of the first flow control mechanism 23 and adjusting the rotational speed of the drive mechanism 21. FIG. 3 shows a schematic diagram of a surge line 30b when the rotational speed changes, in addition to a surge line 30a when the pre-swirl angle changes. As shown by comparing surge lines 30a and 30b, it is generally known that increasing the pre-swirl angle can broaden the stable operating range to a lower airflow rate compared to reducing the rotational speed alone. Therefore, in this embodiment, capacity control is first performed by adjusting the valve opening of the first flow control mechanism 23, and the plant operating point 31 is sufficiently shifted toward the lower airflow rate (the flow rate-discharge pressure characteristic curve of the turbo compressor in FIG. 3 is changed from curve 29a to curve 29d, and the plant flow rate Q is changed from Qspec to Qsw).

[0028] The final step in the capacity control process when using variable speed drive is to reduce the plant flow rate from an operating state where the plant flow rate Q = Qsw to a flow rate point Qc, where the plant flow rate is equal to the minimum stable operating flow rate that the turbo compressor 100 can achieve. When the plant flow rate reaches Qsw, the flow control valve 23a is nearly fully closed, and the flow control valve 23b is nearly fully open. From this state, the rotational speed of the variable speed drive is gradually reduced, changing the flow rate-discharge pressure characteristic curve of the turbo compressor 100 from curve 29d to curve 29e, ultimately achieving an operating state where the plant flow rate Q = Qc.

[0029] When the change in rotation speed relative to the rated rotation speed is small, it can be safely assumed that the change in compressor performance when the rotation speed of the turbo compressor 100 changes will follow the rotation speed similarity law of general turbomachinery (flow rate is proportional to the first power of the rotation speed ratio, head is proportional to the square of the rotation speed ratio, and power is proportional to the cube of the rotation speed ratio).Since the difference between Qsw and Qc is small and the reduction in the compressor rotation speed required to reduce the plant flow rate Q from Qsw to Qc is small, the target reduction in the rotation speed can be derived by a simple calculation based on the rotation speed similarity law.

[0030] Here, the actual method for determining the plant flow rate Qsw can be, for example, as follows. First, for each flow rate ratio between the first suction passage 7 and the second suction passage 8, the suction volume flow rate-polytropic head characteristic curve and operating limit points of the turbo compressor 100 are obtained in advance through experiments and numerical calculations, and a surge line 30a connecting the operating limit points is obtained. With respect to the plant flow rate Qc, which is the intersection of the surge line 30a and the specified discharge pressure Pd,spec when the pre-swirl angle changes, Qsw is determined to be a flow rate larger than this (for example, 5%).

[0031] The effects of this embodiment will be described below. Due to the above-described configuration, the turbo compressor system 1000 of this embodiment can continuously change the flow rate ratio between the gas flow rate passing through the first suction passage 7, which can introduce the gas 25 into the turbo compressor 100 without applying pre-swirl, and the gas flow rate passing through the second suction passage 8, which can introduce the gas 25 into the turbo compressor 100 while applying pre-swirl, simply by controlling the valve openings of the flow control valves 23a and 23b.

[0032] Therefore, it becomes possible to continuously change the pre-swirl angle of the gas 25 introduced into the first stage impeller 1b of the turbo compressor 100, enabling capacity control that continuously changes the operating point of the turbo compressor 100. In this embodiment, because capacity control is performed by applying pre-swirl, capacity control is possible that is more efficient than suction throttling control and has a wider stable operating range on the low air volume side. Furthermore, because continuous capacity control is possible without installing movable inlet guide vanes or their drive mechanisms inside the turbo compressor 100, the compressor structure does not become more complex or larger, and high reliability of the equipment can be ensured.

[0033] In particular, when a fixed speed drive machine is used for the drive machine 21, the capacity can be controlled by adjusting only the first flow rate adjustment mechanism 23, which is advantageous in terms of cost.

[0034] When a variable speed drive is used for the drive machine 21, further improvements in equipment reliability and controllability during capacity control can be expected. In addition, when capacity control is gradually advanced from the high air volume side, including the rated point, to the low air volume side, the pre-swirl angle is first increased by adjusting the valve opening of the first flow control mechanism 23, and then the rotation speed of the drive machine 21 is reduced. This makes it possible to obtain the target amount of rotation speed reduction through simple calculations based on the rotation speed similarity law of turbomachinery.

[0035] Fig. 4 shows a modified example of the turbo compressor system of the first embodiment. As shown in Fig. 4, this modification employs a three-way valve 23c at a branching point 22c where the suction pipe 22 branches into a first suction pipe 22a and a second suction pipe 22b as a first flow rate adjustment mechanism 23 for continuously adjusting the flow rate ratio of gas 25 passing through the first suction passage 7 and the second suction passage 8, thereby configuring a turbo compressor system 1000. The other configurations are the same as those of the first embodiment shown in Fig. 1 described above.

[0036] The following describes the effects of this modification of the present embodiment 1. The configuration of this modification makes it possible to reduce the number of first flow rate adjustment mechanisms 23 installed, thereby reducing costs, and also simplifies valve opening control of the first flow rate adjustment mechanisms 23.

[0037] In a turbo compressor system 1000 installed in a plant, an operator may operate the flow rate adjustment mechanisms during capacity control. If the operator accidentally closes all of the flow rate adjustment mechanisms, the gas 25 will no longer be supplied to the turbo compressor 100. It is known that if the flow rate of the gas 25 falls below the low-flow-side stable operating limit of the turbo compressor 100, the turbo compressor 100 may enter a surging state, in which the gas flow alternates between forward and reverse flow accompanied by large pressure pulsations, potentially damaging the plant. To avoid this, the plant may need to be shut down and restarted, posing a risk of requiring a significant amount of time to restart the plant.

[0038] In this embodiment, a three-way valve 23c is applied to the branching portion 22c of the suction pipe 22 as the first flow rate adjustment mechanism 23. By applying this three-way valve 23c, the first suction pipe 22a and the second suction pipe 22b are prevented from being completely closed at the same time, which makes it possible to reduce the risk of the turbo compressor 100 entering a surging state due to an operator's incorrect operation. [Example]

[0039] A turbo compressor system according to a second embodiment will be described with reference to Fig. 5 and Figs. 2 and 3 which are also used in the description of the first embodiment. Fig. 5 is a diagram for explaining the turbo compressor system according to the second embodiment. In the description of the second embodiment, the description of the same parts as those in the first embodiment will be omitted, and the description will focus on the parts that are different from the first embodiment.

[0040] In the second embodiment shown in FIG. 5 , a bypass line 27 is connected to the second suction pipe 22b downstream of the installation position of a first flow control mechanism 23 (flow control valve 23b) that branches off from the discharge pipe 24 downstream of the turbo compressor 100 and is installed in the second suction pipe 22b. Also, a cooler 26 that adjusts the temperature of gas 25 that passes through the bypass line 27 and flows into the second suction passage 8 and a second flow control mechanism 28 that adjusts the flow rate are installed in the bypass line 27. As in the first embodiment, the first flow control mechanism 23 in this embodiment is configured as flow control valves 23a and 23b, but other types of flow control mechanisms can also be used. As in the modified example of the first embodiment shown in FIG. 4 , the first flow control mechanism 23 may be configured as a three-way valve 23c installed at the branch portion 22c of the suction pipe 22.

[0041] The operation of the turbo compressor system in this embodiment will be described. As shown in Fig. 2, the capacity control method used when gradually reducing the plant flow rate from the rated flow rate Qspec to the flow rate point Qc is the same as in the first embodiment. That is, if the first flow rate adjustment mechanism 23 shown in Fig. 5 is configured as two flow rate adjustment valves 23a and 23b, then during operation at the rated flow rate Qspec, the flow rate adjustment valve 23a is fully open and the flow rate adjustment valve 23b is fully closed.

[0042] From this state, by adjusting the opening of the two flow control valves 23a and 23b, the flow rate of the gas 25 passing through the first suction passage 7 is gradually reduced, while the flow rate of the gas 25 passing through the second suction passage 8 is gradually increased at a rate smaller than the reduction in the flow rate of the gas 25 passing through the first suction passage 7, thereby increasing the pre-swirl angle in the suction gas of the turbo compressor 100 and reducing the plant flow rate.

[0043] Furthermore, when the opening degree of the flow rate adjustment valve is adjusted until the flow rate adjustment valve 23a is fully closed and the flow rate adjustment valve 23b is fully open, operation at a flow rate point Qc where the minimum stable operation flow rate achievable by the turbo compressor 100 alone and the plant flow rate are the same is realized. At this time, the plant operation point 31 coincides with the intersection of the curve 29c, which is the flow rate - discharge pressure characteristic curve of the turbo compressor 100, and the specified discharge pressure Pd,spec.

[0044] Here, depending on the requirements of the plant where the turbo compressor system 1000 is installed, there may be a requirement to operate such that the plant flow rate Q is further reduced (Q < Qc) below the minimum stable operation flow rate achievable by the turbo compressor 100 alone. In such a case, the valve opening degree of the second flow rate adjustment mechanism 28 is gradually increased from the fully closed state, and after adjusting the temperature of the high - temperature gas 25 discharged from the turbo compressor 100 to match the suction gas temperature in the suction pipe 22 by the cooler 26, it is returned through the bypass line 27 into the second suction pipe 22b downstream of the installation position of the first flow rate adjustment mechanism 23.

[0045] By operating in this way, even when the plant flow rate passing through the suction pipe 22 becomes Q < Qc, the flow rate of the gas 25 actually sucked into the turbo compressor 100 through the second suction flow path 8 can be the sum of the plant flow rate Q passing through the suction pipe 22 and the flow rate QB of the gas 25 passing through the bypass line 27.

[0046] At the plant operation point where Q < Qc, by adjusting the opening degree of the second flow rate adjustment mechanism 28 so that QB = Qc - Q, the state where the gas flow rate actually sucked into the turbo compressor 100 does not decrease further can be maintained. Therefore, even when the plant flow rate Q is smaller than the minimum stable operation flow rate achievable by the turbo compressor 100 alone, the compressor can be operated stably.

[0047] Subsequently, as shown in Fig. 3, in this embodiment as well, when gradually reducing the plant flow rate from the rated flow rate Qspec to the flow rate point Qc, the capacity control method is the same as that in the above-described Embodiment 1. That is, first, capacity control by adjusting the valve opening degree of the first flow rate adjustment mechanism 23 is performed first. The flow rate adjustment valve 23a is fully open, and the flow rate adjustment valve 23b is fully closed. While gradually reducing the flow rate of the gas 25 passing through the first suction passage 7 by adjusting the valve opening degree, the flow rate of the gas 25 passing through the second suction passage 8 is gradually increased at a rate smaller than the reduction amount of the flow rate of the gas 25 passing through the first suction passage 7, and the plant flow rate Q is reduced from Qspec to Qsw. When the plant flow rate reaches Qsw, the flow rate adjustment valve 23a is almost fully closed, and the flow rate adjustment valve 23b is fully open. From this state, by gradually reducing the rotational speed of the variable speed drive machine and changing the flow rate - discharge pressure characteristic curve of the turbo compressor 100 from 29d⇒29e, the operating state of the plant flow rate Q = Qc is realized.

[0048] Even when a variable speed drive machine is applied to the drive machine 21, depending on the plant requirements, it may be required to reduce the plant flow rate Q below the minimum stable operating flow rate that the turbo compressor 100 can achieve as a single device, that is, an operation such that Q < Qc. In such a case, while maintaining the rotational speed of the variable speed drive machine the same as the rotational speed in the operating state at Q = Qc, similar to the case where a fixed speed drive machine is applied in Embodiment 2, the valve opening degree of the second flow rate adjustment mechanism 28 is gradually increased from the fully closed state.

[0049] Then, after adjusting the temperature of the high-temperature gas 25 discharged from the turbo compressor 100 to match the suction gas temperature in the suction pipe 22 by the cooler 26, it is returned through the bypass line 27 to the second suction pipe 22b downstream of the installation position of the first flow rate adjustment mechanism 23. At the plant operating point where Q < Qc, by adjusting the opening degree of the second flow rate adjustment mechanism 28 so that the flow rate QB of the gas 25 passing through the bypass line 27 becomes QB = Qc - Q, it becomes possible to stably operate the compressor even when the plant flow rate Q is smaller than the minimum stable operating flow rate that the turbo compressor 100 can achieve as a single device.

[0050] The operation and effects of this embodiment will be described below. Due to the configuration of the turbo compressor system 1000 of this embodiment as described above, first, in addition to obtaining the effects described in Embodiment 1, further, the gas 25 discharged from the turbo compressor 100 with the gas temperature adjusted is returned into the second suction passage 8 through the bypass line 27. Then, if the opening degree of the second flow rate adjustment mechanism 28 is adjusted so that QB = Qc - Q at the plant operating point where Q < Qc, even under the condition that the plant flow rate Q is smaller than the minimum stable operating flow rate that can be achieved by the turbo compressor 100 alone, the turbo compressor 100 can be stably operated. Therefore, it becomes possible to further expand the stable operating range on the low air volume side.

[0051] In addition to this, in this embodiment, the downstream end of the bypass line 27 is connected into the second suction pipe 22b downstream of the installation position of the first flow rate adjustment mechanism 23 installed in the second suction pipe 22b. Thereby, even under the condition that the plant flow rate Q is smaller than the minimum stable operating flow rate that can be achieved by the turbo compressor 100 alone, all the gas 25 sucked into the turbo compressor 100, including the gas 25 recirculated to the suction pipe 22 side through the bypass passage, is given pre-rotation and introduced into the compressor. Therefore, even at the low flow rate side plant operating point where Q < Qc, capacity control capable of reducing the axial power on the high efficiency and low air volume side can be performed.

[0052] In this embodiment, an example was described in which the valve opening degree of the second flow rate adjustment mechanism 28 provided in the bypass line 27 is gradually increased from the fully closed state at a plant flow rate where Q < Qc. However, in order to further reduce the risk that the turbo compressor 100 will enter surging, the valve opening degree of the second flow rate adjustment mechanism 28 may be gradually increased from a plant operating point that is slightly on the high flow rate side with respect to the surge line 30a during the pre-whirl angle change. As a line for starting the adjustment of the valve opening degree of the second flow rate adjustment mechanism 28 in this case, for example, a line obtained by parallelly shifting the surge line 30a during the pre-whirl angle change shown in FIG. 2 or FIG. 3 to the right side of the graph by, for example, 5% is set, and when the plant flow rate Q reaches this line, the valve opening degree of the second flow rate adjustment mechanism 28 may be operated so as to gradually increase the valve opening degree of the second flow rate adjustment mechanism 28.

Embodiment

[0053] Embodiment 3 will be described with reference to FIGS. 6 to 9. FIG. 6 is a diagram for explaining the turbo compressor used in the turbo compressor system of Embodiment 3, and is a longitudinal sectional view of a single-shaft multi-stage centrifugal compressor. Further, FIG. 7 shows a view taken in the direction of arrow I-I in FIG. 6, FIG. 8 shows a view taken in the direction of arrow II-II in FIG. 6, and FIG. 9 shows a view taken in the direction of arrow III-III in FIG. 6.

[0054] In this Embodiment 3, in the turbo compressor system 1000 described in Embodiment 1 and Embodiment, the turbo compressor 100 is configured as a single-shaft multi-stage centrifugal compressor 100a. The single-shaft multi-stage centrifugal compressor 100a of this embodiment has a plurality of impellers 1, which are centrifugal impellers, attached to a single rotating shaft 2 in multiple stages. Both end portions of the rotating shaft 2 are rotatably supported by bearings 3b provided in a bearing case 3a. The casing 4 is cylindrical and houses a plurality of impellers 1. The casing 4 is provided with an inlet nozzle 5 for introducing gas 25 into the first-stage impeller 1b among the plurality of impellers 1, and a discharge nozzle 6 provided in the casing 4 for discharging the fluid discharged from the final-stage centrifugal impeller among the plurality of impellers 1.

[0055] The introduction nozzle 5 is composed of a first introduction nozzle 5a and a second introduction nozzle 5b. The introduction nozzle 5 is also composed of an inlet flange portion 5aa of the first introduction nozzle 5a and an inlet flange portion 5ba of the second introduction nozzle 5b. The first suction passage 7 is a passage that allows the fluid sucked through the first introduction nozzle 5a to flow into the first-stage impeller 1b. The second suction passage 8 is a passage that allows the fluid sucked through the second introduction nozzle 5b to flow into the first-stage impeller 1b.

[0056] The first suction passage 7 includes a nozzle passage section 7a, which is a passage formed within the first introduction nozzle 5a, an annular passage section 7b that is provided in a ring shape around the rotating shaft 2 and that causes the working fluid to flow around in the circumferential direction, and an L-shaped bend passage section 7c that is provided on the inner peripheral side of the annular passage section 7b and that redirects the flow that has passed through the annular passage section 7b from the radially inward direction to the downstream direction of the rotating shaft 2 and leads it to the first-stage impeller 1b.

[0057] Similarly, the second suction passage 8 also includes a nozzle passage section 8a, which is a passage formed within the second introduction nozzle 5b, an annular passage section 8b that is provided in a ring shape around the rotating shaft 2 and causes the working fluid to flow around in the circumferential direction, and an L-shaped bend passage section 8c that is provided on the inner peripheral side of the annular passage section 8b and redirects the flow that has passed through the annular passage section 8b from the radially inward direction to the downstream direction of the rotating shaft 2 and leads it to the first-stage impeller 1b.

[0058] The second suction passage 8 is configured to impart a swirling component to the flow that flows into the first-stage impeller 1b, while the first suction passage 7 is configured to allow a flow that does not impart a swirling component to flow into the first-stage impeller 1b.

[0059] A diffuser section 9, which serves as a flow path for the working gas discharged from each stage impeller, is formed in the radial direction on the radially outer side of the impeller 1 of each stage. In this embodiment, the diffuser section 9 is configured as a vaned diffuser having a plurality of blades 9a arranged at intervals in the circumferential direction. Note that a vaneless diffuser, which does not have any blades at all, may also be used for the diffuser section 9.

[0060] A return channel 10 is provided downstream of the diffuser section 9 to change the flow of working gas from a radially outward flow to a radially inward flow in order to form a suction flow path to the next-stage impeller 1. Return vanes 10a are arranged at intervals in the circumferential direction in this return channel 10 to straighten the flow of working gas. The diffuser section 9 and the return channel 10 form a stationary flow path 11.

[0061] A scroll 12 is formed downstream of the final stage impeller 1 and radially outward, and is configured to collect the high-pressure working gas flowing out from the final stage impeller 1 and discharge it from the discharge nozzle 6 to the outside of the machine.

[0062] A shaft seal 13 is provided to seal the gap between the rotating shaft 2 and the casing 4, between the diffuser section 9 and the return channel 10. In this embodiment, this shaft seal 13 is configured as a labyrinth seal, which is attached to the inner circumferential surface of the shaft-through portion of the casing 4 and is disposed with a small gap between it and the rotating shaft 2. This shaft seal 13 prevents the working gas discharged from the impeller 1 in the previous stage from flowing into the impeller 1 in the subsequent stage through the gap.

[0063] Furthermore, shaft seals 14 and 15 each made of a labyrinth seal are provided at the shaft through-hole between the first-stage impeller 1 and bearing 3b and at the shaft through-hole between the final-stage impeller 1 and bearing 3b. Shaft seal 14 prevents air from outside the machine from flowing into the first-stage impeller 1, and shaft seal 15 prevents the working gas discharged from the final-stage impeller 1 from leaking directly outside the machine without flowing into the scroll 12.

[0064] The bearing case 3a is provided integrally with the casing 4, or a mounting base is provided separately to mount the bearing case 3a.

[0065] An intake passage is formed at the end of the cylindrical casing 4, which houses the plurality of impellers 1 and the rotating shaft 2, on the side of the first-stage impeller 1b, and in this embodiment, the intake passage is made up of the above-mentioned first intake passage 7 and the second suction passage 8, which is installed closer to the first-stage impeller 1 than the first suction passage 7. These first suction passage 7 and second suction passage 8 are passages for introducing working gas from the outside of the single-shaft multi-stage centrifugal compressor 100a into the compressor, and are configured to convert the flow flowing into the first-stage impeller 1 from the radial direction into an axial intake flow.

[0066] A first introduction nozzle 5a and a second introduction nozzle 5b are provided on the cylindrical outer peripheral surface of the suction side of the casing 4, and their inlets are inlet flanges 5aa and 5ba. The cross-sectional shapes of these inlet flanges 5aa and 5ba are circular. The cross-sectional shapes of the first introduction nozzle 5a and the second introduction nozzle 5b are configured to become oval (becoming increasingly flat) that are plane-symmetrical with respect to the center cross section of the compressor shown in Figure 1 as they move toward the inner diameter side.

[0067] The working gas sucked through the first introduction nozzle 5a or the second introduction nozzle 5b is guided radially inward via the first introduction passage or the second introduction passage and flows into the suction port 1a of the first-stage impeller 1b. Here, the passage from the inlet flange portion 5aa of the first introduction nozzle 5a to the suction port 1a of the first-stage impeller 1 is the first introduction passage 7, and the passage from the inlet flange portion 5ba of the second introduction nozzle 5b to the suction port 1a of the first-stage impeller 1 is the second introduction passage 8.

[0068] As described above, the first suction passage 7 and the second suction passage 8 are composed of the nozzle passage section 7a, the nozzle passage section 8a, the annular passage section 7b, the annular passage section 8b, the L-shaped bend passage section 7c, and the L-shaped bend passage section 8c. The annular passage section 7b and the annular passage section 8b are connected to the inner diameter sides of the nozzle passage section 7a and the nozzle passage section 8a, and direct the flow introduced from the first introduction nozzle 5a and the second introduction nozzle 5b in the circumferential direction. The L-shaped bend passage section 7c and the L-shaped bend passage section 8c are connected to the inner diameter sides 7ba and 8ba of the annular passage section 7b and the annular passage section 8b, and direct the working gas from the radially inward direction to the downstream direction of the axis X of the rotating shaft 2.

[0069] A plurality of fixed fluid guides 17 are installed in an annular arrangement at intervals in the circumferential direction in the L-shaped bend flow path portion 8c of the second suction flow path 8. These fixed fluid guides 17 are intended to give a stable and uniform swirling flow to the fluid flowing through the second suction flow path 8.

[0070] In addition, the first suction passage 7 and the second suction passage 8 are separated by a partition wall 18 that forms part of the casing 4, and the L-shaped bend passage section 7c and the L-shaped bend passage section 8c merge downstream of the inner diameter side tip portion 18a of the partition wall 18 and are configured to connect to the suction port 1a of the first stage impeller 1b.

[0071] Other elements constituting the turbo compressor system 1000 in this embodiment, as well as the operating method, are the same as those shown in the first and second embodiments.

[0072] Next, the configurations of the inlet nozzles 5 (first inlet nozzle 5a, second inlet nozzle 5b) and the suction passages (first suction passage 7, second suction passage 8) for introducing the working gas into the first stage impeller 1b in the single-shaft multi-stage centrifugal compressor 100a of this embodiment, and the flow state of the working gas passing through the suction passages will be described with reference to FIGS. 7 to 9.

[0073] FIG. 7 is a view taken along line II in FIG. 6 , showing the single-shaft multi-stage centrifugal compressor 100a as viewed from the upstream side in the direction of the axis X of the rotating shaft 2. An example of the arrangement positions of the first inlet nozzle 5a and the second inlet nozzle 5b will be described with reference to FIG. 7 . As shown in FIG. 7 , the first inlet nozzle 5a constituting the first suction passage 7 in this embodiment is arranged facing the direction of the rotating shaft 2. On the other hand, the second inlet nozzle 5b constituting the second suction passage 8 is arranged facing in a direction offset from the direction of the rotating shaft 2. Furthermore, the second inlet nozzle 5b constituting the second suction passage 8 is installed in a position preceding the first inlet nozzle 5a constituting the first suction passage 7 in the direction of rotation A of the impeller 1 and the rotating shaft 2, as viewed from the upstream direction of the rotating shaft 2.

[0074] Fig. 8 is a cross-sectional view taken along line II-II in Fig. 6, showing the shape of the first suction passage 7 and the flow direction of the working gas in Example 3. Fig. 9 is a cross-sectional view taken along line III-III in Fig. 6, showing the shape of the second suction passage 8 and the flow direction of the working gas.

[0075] As shown in Figure 8, in this embodiment, the annular flow path portion 7b of the first suction flow path 7 is configured so that the flow path cross-sectional shape is a substantially constant shape along the circumferential direction of the rotating shaft 2. In addition to the flow path shape of the first suction flow path 7, Figure 8 also shows the flow of the working gas when only the flow rate adjustment valve 23a shown in Figures 1, 5, and 6 is open and the flow rate adjustment valve 23b is closed.

[0076] The flow (flow direction) of the working gas introduced from the inlet flange 5aa of the first introduction nozzle 5a into the first suction passage 7 is indicated by arrow B. The working gas introduced from the inlet flange 5aa of the first introduction nozzle 5a into the first suction passage 7 passes through the nozzle passage 7a and flows into the annular passage 7b. The working gas flows around the entire circumference of the annular passage 7b, and as indicated by arrow B, the flow direction at each circumferential position is directed toward the axis X of the rotating shaft 2, i.e., a flow direction without a swirling component, forming a symmetrical flow field before flowing into the L-shaped bend passage 7c. The working gas that has flowed into the L-shaped bend passage 7c changes its flow direction downstream of the axis X and flows toward the suction port 1a of the first-stage impeller 1b.

[0077] In this embodiment, as shown in FIG. 8, a swirl prevention plate 19 is provided inside the first suction passage 7 on the opposite side of the position where the first introduction nozzle 5a is installed, from the annular passage section 7b to the L-shaped bend passage section 7c, in order to eliminate swirl in the flow of working gas and ensure symmetry of the flow.

[0078] In this way, the first suction passage 7 is a suction passage that introduces the working gas into the suction port 1a of the first-stage impeller 1 without any swirling component.

[0079] 9, the annular flow passage portion 8b in the second suction flow passage 8 is configured so that its flow passage cross-sectional shape is substantially constant along the circumferential direction of the rotating shaft 2. However, the flow passage cross-sectional shape of the annular flow passage portion 8b in the second suction flow passage 8 is not limited to this, and may be configured, for example, so that it has a spiral shape in which the flow passage cross-sectional area gradually narrows along the circumferential direction of the rotating shaft 2 from the junction C of the nozzle flow passage portion 8a of the second introduction nozzle 5b and the annular flow passage portion 8b. In addition to the flow passage shape of the second suction flow passage 8, FIG. 9 also shows the flow of the working gas when only the flow rate control valve 23b shown in FIGS. 1, 5, and 6 is open and the flow rate control valve 23a is closed.

[0080] The flow (flow direction) of the working gas introduced from the inlet flange portion 5ba of the second introduction nozzle 5b into the second suction passage 8 is also indicated in this figure by arrow B. The working gas introduced from the inlet flange portion 5aa of the second introduction nozzle 5b into the first suction passage 7 passes through the nozzle passage portion 7a and flows into the annular passage portion 8b.

[0081] The working gas introduced into the second suction passage 8 from the inlet flange portion 5ba of the second inlet nozzle 5b passes through the nozzle passage portion 8a and then flows into the annular passage portion 8b, as shown by arrow B. In this embodiment, the second inlet nozzle 5b is installed in a direction offset from the direction of the axis X of the rotating shaft 2 and in a direction preceding the first inlet nozzle 5a in the rotation direction A of the rotating shaft 2. However, this is not necessarily limited to this, and the first inlet nozzle 5a constituting the first suction passage 7 and the second inlet nozzle 5b constituting the second suction passage 8 may both be arranged facing the direction of the rotating shaft 2.

[0082] The working gas that flows into the annular flow passage 8b gradually flows inward through the annular flow passage 8b while swirling in a fixed direction, i.e., the same direction as the rotational direction A of the rotating shaft 2 and the impeller 1 in FIG. 9 , and then flows into the L-shaped bend flow passage 8c. Furthermore, in the L-shaped bend flow passage 8c, multiple fixed fluid guides 17 are annularly arranged at regular intervals in the circumferential direction around the axis X as the center of rotation. Therefore, the fixed fluid guides 17 straighten the working gas to achieve a desired swirl angle. While the fixed fluid guides 17 are configured as straight blades with a uniform blade thickness in FIG. 9 , they can also be configured as blades with a curved cross-sectional shape with a blade thickness distribution. Alternatively, the shape of the fixed fluid guides 17 may be varied circumferentially to increase the uniformity of the swirl angle distribution of the flow.

[0083] The working gas that has flowed into the L-shaped bend flow path portion 8c turns its flow direction toward the downstream side of the axis X while swirling, and flows toward the suction port 1a of the first-stage impeller 1b. In this way, the second suction path 8 can introduce the working gas into the suction port 1a of the first-stage impeller 1b by imparting a swirling flow that has been rectified to have a desired swirl angle to the flow of the working gas by the fixed fluid guide 17.

[0084] In this third embodiment, as shown in Figure 9, a plurality of fixed fluid guides 17 are annularly arranged at regular intervals in the circumferential direction of the L-shaped bend passage section 8c, with the axis X serving as the center of rotation, as a structure for imparting pre-swirl to the gas. By providing these fixed fluid guides 17 for imparting pre-swirl, the flow resistance of the gas is greater in the second suction passage 8 than in the first suction passage 7. Therefore, in order to gradually reduce the plant flow rate from the rated flow rate Qspec, the openings of the two flow control valves 23a and 23b can be adjusted as follows.

[0085] At the rated flow rate Q = Qspec, the compressor is operated with the flow control valve 23a fully open and the flow control valve 23b fully closed. To reduce the gas flow rate, the flow control valve 23b is also fully opened. Even if both flow control valves 23a and 23b are fully opened, the difference in flow resistance causes the gas to selectively pass through the first suction passage 7. This prevents a sudden increase in the pre-swirl angle of the gas introduced into the compressor, and the plant operating point does not suddenly change from Qspec. After the flow control valve 23b is fully opened, the pre-swirl angle is further increased by gradually reducing the valve opening of the flow control valve 23a. The flow rate is reduced to the flow point Qc, where the plant flow rate is equal to the minimum stable operating flow rate achievable by the single-shaft multi-stage centrifugal compressor 100a.

[0086] Next, the effects of this embodiment will be described. The turbo compressor system 1000 of this embodiment is configured as a single-shaft multi-stage centrifugal compressor 100a, in which both ends of a rotating shaft 2 are supported by bearings 3b housed in bearing cases 3a, as the turbo compressor 100 installed inside the system. In this type of single-shaft multi-stage centrifugal compressor 100a, as shown in Figures 6 to 9, a system must be adopted in which gas 25 is taken in through an inlet nozzle 5 installed in a radially outward direction, and as it flows down through the first suction passage 7 and the second suction passage 8, the flow direction is changed from the radial direction to the direction of the rotating shaft, and the gas is sucked into the suction port 1a of the first-stage impeller 1b.

[0087] Therefore, the bearing 3b that supports the rotating shaft 2 installed on the suction passage side needs to be installed outside the casing 4, with the suction passage forming portion sandwiched between them, and the axial distance (bearing span) between the bearings 3b tends to be large. Here, in the single-shaft multi-stage centrifugal compressor 100a, the number of stages of the impeller 1 mounted on the rotating shaft 2 is often large, so the bearing span tends to be large. In addition to these characteristics, if the axial length of the suction passage portion is further increased, it will lead to an increase in the axial length and bearing span of the rotating shaft 2, which will reduce the rigidity of the rotating shaft 2 and lower the critical speed of the shaft system, making it impossible to sufficiently increase the rotation speed of the compressor.

[0088] When a movable inlet guide vane having a drive mechanism is installed as a capacity control means for the single-shaft multi-stage centrifugal compressor 100a, the installation of the drive mechanism increases the axial length of the suction passage, which may prevent the rated rotation speed of the compressor from being increased sufficiently and may prevent the compressor from being downsized. However, by adopting the configuration of this embodiment, continuous capacity control of the single-shaft multi-stage centrifugal compressor 100a is possible without installing a movable inlet guide vane or its drive mechanism, and it is possible to suppress an increase in the bearing span of the single-shaft multi-stage centrifugal compressor 100a.

[0089] Therefore, in addition to the effects obtained in Examples 1 and 2, this example enables highly reliable capacity control over a wide operating range with high efficiency without reducing the rated rotation speed of the single-shaft multi-stage centrifugal compressor 100a.

[0090] Next, Fig. 10 shows a modified turbo compressor system of the third embodiment of the present invention. As shown in Fig. 10, this modification employs only a flow rate adjustment valve 23a installed in the first suction pipe 22a as the first flow rate adjustment mechanism 23 for continuously adjusting the flow rate ratio of the gas passing through the first suction passage 7 and the second suction passage 8, thereby configuring a turbo compressor system 1000. The other configurations are the same as those of the third embodiment shown in Figs. 6 to 9 described above.

[0091] In Example 3, a fixed fluid guide 17 for imparting pre-swirl is installed as shown in Figure 9, and the flow path resistance of gas is greater in the second suction path 8 than in the first suction path 7. As a result, when the flow control valve 23a installed in the first suction pipe 22a is fully opened, the difference in flow path resistance causes gas to pass almost exclusively through the first suction path 7, even if the flow control valve 23b, which is the first flow control mechanism, is not installed in the second suction pipe 22b. Therefore, the ratio of the gas flow rates through the first suction path 7 and the second suction path 8 can be adjusted simply by adjusting the opening of the flow control valve 23a installed in the first suction pipe 22a.

[0092] The effects of the modification of the present embodiment 3 will be described. In addition to the effects obtained in the embodiments 1 to 3, the modification of the present embodiment 3 can further simplify the arrangement, number, and installation configuration of the flow control valves, thereby achieving further cost reduction and improved valve operability. This configuration further reduces the risk of both suction flow paths being closed simultaneously, and can also reduce the risk of compressor surging. In addition, in this modification, there is no pressure loss in the second suction pipe 22b when the valve opening of the flow control valve 23b in the second suction pipe 22b is smaller than in the fully open state. This also provides the effect of improving the efficiency of the compressor in a transient state during capacity control from the rated flow point to the flow point Qc at which the plant flow rate is equal to the minimum stable operating flow rate that can be achieved by the single-shaft multistage centrifugal compressor 100a. [Example]

[0093] A turbo compressor system according to a fourth embodiment will be described with reference to Figures 11 to 14. The fourth embodiment describes an example of the turbo compressor system according to the first to third embodiments, in which the turbo compressors used in the turbo compressor system are configured in multiple groups or in which the turbo compressor system is configured with multiple turbo compressors. Note that a description of parts having the same configuration as those in the first to third embodiments will be omitted.

[0094] FIG. 11 is a diagram showing an example of a turbo compressor system according to a fourth embodiment, in which one turbo compressor is configured with two groups, a first group and a second group. FIG. 12 is another example, in which two turbo compressors (a low-pressure compressor 100L and a high-pressure compressor 100H) are configured with two groups. FIG. 13 is a diagram illustrating an example of a turbo compressor system used in a synthesis plant for ammonia, methanol, or the like, in which the turbo compressor is configured with two turbo compressors: a low-pressure compressor 100L configured with a first group and a second group, and a high-pressure compressor 100H configured with a third group and a recycle stage. FIG. 14 is a diagram illustrating another example of a turbo compressor system according to the fourth embodiment, in which the turbo compressor is configured with one turbo compressor configured with a first group, a second group, and a recycle stage.

[0095] Here, each group of turbo compressors is defined as the region from the suction passage to the discharge passage, through which gas is introduced from outside the turbo compressor, pressurized inside the compressor, and then discharged outside. Group 1 indicates the group of turbo compressors that is furthest upstream, and gas circulating in the plant first passes through this group to be pressurized, and then is further pressurized in the second and third groups, before being discharged at the final discharge pressure.

[0096] In the turbo compressor system 1000 shown in FIG. 11 , the turbo compressor 100 is configured as a single compressor composed of a first group and a second group. In this embodiment, the suction pipes 22 connected to the first and second groups of the turbo compressor 100 are configured as a first suction pipe 22a that introduces gas 25 into the compressor without pre-swirl and a second suction pipe 22b that introduces gas 25 into the compressor with pre-swirl. Furthermore, a three-way valve 23c installed at the pipe branch is used as a first flow control mechanism 23 for adjusting the flow rate ratio between the first suction pipe 22a and the second suction pipe 22b. A bypass line 27 branches off from the discharge pipe 24 of the second group and connects to the second suction pipe 22b of the first group.

[0097] In the turbo compressor system 1000 shown in Figure 12, the turbo compressor 100 is configured with two turbo compressors: a low-pressure compressor 100L made up of two groups, a first group and a second group, and a high-pressure compressor 100H made up of two groups, a third group and a fourth group. Both the low-pressure compressor 100L and the high-pressure compressor 100H have individual configurations similar to those of the turbo compressor 100 shown in Figure 11. A cooler 26 is installed in the discharge piping 24 of the second group, and a piping branch is provided downstream of this cooler 26 to a bypass line 27 connected to the second suction piping 22b of the first group and to the suction piping 22 of the third group of the high-pressure compressor 100H.

[0098] The turbo compressor system 1000 shown in Figure 13 is an example used in a synthesis plant for ammonia, methanol, etc. The turbo compressor 100 is composed of two turbo compressors: a low-pressure compressor 100L composed of a first group and a second group, and a high-pressure compressor 100H composed of a third group and a recycle stage.

[0099] In a synthesis plant for ammonia, methanol, etc., most of the pressure increase to the final discharge pressure required for the synthesis reaction to obtain the final products, ammonia and methanol, from a gas that is a mixture of multiple gases is generally completed by the time it reaches the recycle stage compressor (in Figure 13, the discharge piping 24 of the third group of the high-pressure compressor 100H), and the gas is further pressurized slightly in the recycle stage compressor to reach the final discharge pressure before being discharged to the synthesis reactor. Most of the gas pumped to the synthesis reactor becomes the final product gas, such as ammonia or methanol, but some remains as unreacted gas, which is circulated back to the suction piping 22 of the recycle stage compressor (referred to as recycle gas 32), mixed with gas 25 discharged from the discharge piping 24 of the third group, pressurized by the recycle stage compressor, and discharged again to the synthesis reactor.

[0100] As mentioned above, the amount of gas pressure increase required for the recycle stage compressor is small. Therefore, in the recycle stage compressor, flow separation, stall, and the like are less likely to occur in the internal flow paths of the compressor, such as the impeller flow path, on the low flow rate side, and the stable operation range on the low flow rate side is wider than that of the first to third group compressors. Therefore, the embodiment shown in Figure 13 differs from the turbo compressor 100 shown in Figure 12, which has a similar configuration, in that the suction piping 22 of the recycle stage compressor is composed of only the first suction piping 22a, which introduces gas into the compressor without pre-swirl.

[0101] 14 shows another example of a turbo compressor system 1000 when used in a synthesis plant for ammonia, methanol, etc., similar to the turbo compressor system 1000 shown in Fig. 13. In this example, the turbo compressor 100 is configured from a single turbo compressor made up of a first group, a second group, and a recycle stage.

[0102] The present invention is not limited to the above-described embodiment, but includes various modifications. Furthermore, the present invention is not limited to synthesis gas compressors used in ammonia, methanol, and other synthesis plants, and can be applied to compressors in various other plants.

[0103] The above-described embodiments have been described in detail to clearly explain the present invention, and the present invention is not necessarily limited to those including all of the described configurations. Furthermore, it is possible to add, delete, or replace part of the configuration of each embodiment with other configurations. [Explanation of symbols]

[0104] 1... impeller, 1a... suction port, 1b... first stage impeller, 2...rotating shaft, 3a...bearing case, 3b...bearing, 4...casing, 5...inlet nozzle, 5a...first inlet nozzle, 5aa...inlet flange portion, 5b...second introduction nozzle, 5ba...inlet flange portion, 6...discharge nozzle, 7...first suction flow path, 7a...nozzle flow path portion, 7b...annular flow path portion, 7ba...inner diameter side, 7c...L-shaped bend flow path portion, 8... second suction flow path, 8a... nozzle flow path portion, 8b... annular flow path portion, 8ba...inner diameter side, 8c...L-shaped bend flow path portion, 9...diffuser portion, 9a...wing, 10...return channel, 10a...return vane, 11... stationary flow path, 12... scroll, 13... shaft seal portion, 14... shaft seal portion, 15... shaft seal portion, 17... fixed fluid guide, 18... partition wall, 18a... inner diameter side tip portion, 19... anti-swirl plate, 21... driver, 22... suction pipe, 22a... first suction pipe, 22b... second suction pipe, 22c...Branch of the first and second suction pipes, 23...first flow rate adjustment mechanism, 23a...flow rate adjustment valve, 23b...flow rate adjustment valve, 23c... Three-way valve, 24... Discharge piping, 25... Gas, 26... Cooler, 27...bypass line, 28...second flow rate adjusting mechanism, 29...curve, 30...surge line, 30a...surge line, 30b...surge line, 31... plant operating point, 32... recycled gas, 40... flow rate control device, A: Impeller rotation direction, B: Arrow indicating the flow of working gas, C: Confluence, X: Axis

Claims

1. a turbo compressor having a driver, a rotary shaft that transmits rotation of the driver, and an impeller that is attached to the rotary shaft and compresses gas; and a suction pipe that supplies gas to the turbo compressor, the suction pipe branching into a first suction pipe and a second suction pipe upstream of the turbo compressor, an intake passage that introduces gas into a first-stage impeller of the turbo compressor includes a first intake passage connected to the first intake pipe that introduces the gas into the turbo compressor without imparting swirl to the gas, and a second intake passage connected to the second suction pipe that imparts swirl to the gas and introduces it into the turbo compressor, a turbo compressor system comprising: at least one first flow rate adjustment mechanism that can continuously adjust a flow rate ratio of gas passing through the first suction passage and the second suction passage of the turbo compressor;

2. 2. The turbocompressor system of claim 1, The turbo compressor system according to claim 1, wherein the first flow rate adjustment mechanism is a flow rate adjustment valve individually installed in the first suction pipe and the second suction pipe.

3. 2. The turbocompressor system of claim 1, The turbo compressor system according to claim 1, wherein the first flow rate adjustment mechanism is a three-way valve installed at a branch point of the first suction pipe and the second suction pipe.

4. 2. The turbocompressor system of claim 1, a bypass line branching from a discharge piping downstream of the turbo compressor and connected to the second suction passage downstream of an installation position of the first flow control mechanism; and a second flow control mechanism installed on the bypass line and adjusting a flow rate of fluid that passes through the bypass line and flows into the second suction passage while being temperature-controlled.

5. 2. The turbocompressor system of claim 1, the turbo compressor is a single-shaft multi-stage centrifugal compressor in which both ends of the rotary shaft are supported by bearings, a plurality of centrifugal impellers are attached, and the suction flow path introduces gas from the radially outer side of the rotary shaft and turns the gas in the axial direction while causing it to flow into the first-stage impeller, the gas pre-swirl imparting structure installed in the second suction passage is a plurality of fixed fluid guides arranged in an annular shape at intervals in the circumferential direction, the fixed fluid guides imparting a swirl component to the flow.

6. 6. The turbocompressor system of claim 5, The turbo compressor system according to claim 1, wherein the first flow rate adjustment mechanism is a flow rate adjustment valve individually installed in the first suction pipe and the second suction pipe.

7. 6. The turbocompressor system of claim 5, The turbo compressor system according to claim 1, wherein the first flow rate adjustment mechanism is a three-way valve installed at a branch point of the first suction pipe and the second suction pipe.

8. 6. The turbocompressor system of claim 5, a first flow rate adjusting mechanism that adjusts the flow rate of the first suction pipe; a second flow rate adjusting valve that adjusts the flow rate of the first suction pipe;

9. 6. The turbocompressor system of claim 5, a bypass line branching from a discharge piping downstream of the turbo compressor and connected to the second suction passage downstream of an installation position of the first flow control mechanism; and a second flow control mechanism installed on the bypass line and adjusting a flow rate of fluid that passes through the bypass line and flows into the second suction passage while being temperature-controlled.

10. 5. A method for operating a turbo-compressor system according to claim 4, comprising: the drive is a fixed speed drive; a method for operating a turbo compressor system, the method comprising: when reducing an intake gas flow rate from a rated gas flow rate, starting from an operating state corresponding to a rated operation state in which a valve aperture of the second flow control mechanism is fully closed, and a flow rate ratio between the first intake passage and the second intake passage is maximum and the gas flow rate passing through the first intake passage is greater, adjusting the valve aperture of the first flow control mechanism to gradually reduce the gas flow rate of the first intake passage while increasing the gas flow rate of the second intake passage, and then fully closing the first intake passage and fully opening the second intake passage to apply pre-swirl and thereby reducing the intake gas flow rate passing through the suction piping to a minimum stable operating point that can be achieved by the turbo compressor.

11. 5. A method for operating a turbo-compressor system according to claim 4, comprising: the drive is a fixed speed drive; a method for operating a turbo compressor system, in which, when reducing an intake gas flow rate from a rated gas flow rate, starting from an operating state corresponding to a rated operation state in which a valve aperture of the second flow control mechanism is fully closed, a flow rate ratio between the first intake passage and the second intake passage is maximum, and the gas flow rate passing through the first intake passage is greater, the method first adjusts the valve aperture of the first flow control mechanism to gradually reduce the gas flow rate of the first intake passage while increasing the gas flow rate of the second intake passage, then fully closes the first intake passage, fully opens the second intake passage, and applies pre-swirl to reduce the intake gas flow rate to a minimum stable operating point that can be achieved by the turbo compressor, and then gradually opens the aperture of the second flow control mechanism to gradually increase the gas flow rate passing through the bypass line and flowing into the second intake passage, thereby reducing the intake gas flow rate passing through the suction piping.

12. 5. A method for operating a turbo-compressor system according to claim 4, comprising: the drive is a variable speed drive, a method for operating a turbo compressor system, in which, when reducing the suction gas flow rate from a rated gas flow rate, from an operating state corresponding to a rated operation state in which the valve aperture of the second flow control mechanism is fully closed and the flow rate ratio between the first suction passage and the second suction passage is maximum and the gas flow rate passing through the first suction passage is greater, the valve aperture of the first flow control mechanism is adjusted to gradually reduce the gas flow rate of the first suction passage while increasing the gas flow rate of the second suction passage, until the first suction passage is almost fully closed and the second suction passage is fully opened, thereby further reducing the suction gas flow rate by applying pre-swirl, and then the rotation speed of the driver is reduced to reduce the suction gas flow rate passing through the suction piping to a minimum stable operating point that can be achieved by the turbo compressor.

13. 5. A method for operating a turbo-compressor system according to claim 4, comprising: the drive is a variable speed drive, a turbo compressor system operation method, in which, when reducing an intake gas flow rate from a rated gas flow rate, a valve aperture of the second flow control mechanism is fully closed, which corresponds to a rated operation state, and an operating state in which a flow rate ratio between the first intake passage and the second intake passage is maximum and the gas flow rate passing through the first intake passage is greater, the valve aperture of the first flow control mechanism is adjusted to gradually reduce the gas flow rate of the first intake passage while increasing the gas flow rate of the second intake passage, the first intake passage is almost fully closed, the second intake passage is fully opened, and the intake gas flow rate is further reduced by applying a pre-swirl, and then the rotation speed of the driver is reduced to reduce the intake gas flow rate to a minimum stable operating point that can be achieved by the turbo compressor, and the aperture of the second flow control mechanism is gradually opened to gradually increase the gas flow rate passing through the bypass line and flowing into the second intake passage, thereby reducing the intake gas flow rate passing through the suction piping.

14. The method for operating a turbo compressor system according to any one of claims 10 to 13, a suction volume flow rate-polytropic head characteristic curve and an operating limit point of the turbo compressor are determined in advance for each flow rate ratio of the first suction passage and the second suction passage, a surge line is derived by connecting these curves, and a minimum stable operating point that can be achieved by the turbo compressor is determined.

15. 2. The turbocompressor system of claim 1, A turbocompressor system, wherein the turbocompressor is a single-shaft multi-stage centrifugal compressor for synthesis of ammonia or methanol.

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