US20160327049A1

US20160327049A1 - Multi-stage compression system and method of operating the same

Info

Publication number: US20160327049A1
Application number: US15/215,213
Authority: US
Inventors: Britton D. Dinsdale; Michael O. Muller; James L. Robb, IV
Original assignee: Ingersoll Rand Co
Current assignee: Ingersoll Rand Industrial US Inc
Priority date: 2006-02-13
Filing date: 2016-07-20
Publication date: 2016-11-10
Also published as: CN101421519B; CN101421519A; US20070189905A1; EP1984628B1; EP1984628A4; WO2007095537A1; EP1984628A1

Abstract

A multi-stage compression system includes a plurality of centrifugal compression stages. Each stage includes an impeller and a variable speed motor coup led to the impeller. Each variable speed motor is operable at a speed between a (test speed and a second speed. The multi-stage compression system also includes a control system that is connected to each of the variable speed motors and is operable to vary the speed of each motor. The speed of each motor is varied simultaneously such that a ratio of the speed of any two variable speed motors remains constant.

Description

RELATED APPLICATION DATA

This application claims benefit under 35 U.S.C. Section 119(e) of co-pending U.S. Provisional Application No. 60/772,715, filed Feb. 13, 2006, which is fully incorporated herein by reference.

BACKGROUND

The present invention relates to a system and method to control a centrifugal compression system. In particular, the invention relates to a control system and method that varies the speed of multiple compression stages while maintaining constant speed ratios.
Compressors, and more particularly centrifugal compressors, operate across a wide range of operating parameters. Variation of some of these parameters may produce undesirable efficiency and capacity variations.
Compression of a gas in centrifugal compressors, also known as dynamic compressors, is based on the transfer of energy from a set of rotating blades to a gas such as air. The rotating blades impart energy by changing the momentum and pressure of the gas. The gas momentum, which is related to kinetic energy, is then converted into pressure energy by decreasing the velocity of the gas in stationary diffusers and downstream collecting systems.
The performance of a multi-stage centrifugal compression system is impacted by conditions of the gas al the inlet, such as pressure, temperature, and relative humidity, and by the operating speed of the compressor stages. Specifically, for a given rotational speed of a stage, variations in pressure, temperature, and relative humidity of the gas at the inlet of that compression stage will alter the compressor discharge head and capacity. Additionally, a change in the operating rotational speed of a stage of compression also results in a change in the performance parameters of the overall compressor in terms of discharge head, capacity, and thermodynamic efficiency. It is relevant to note that in dynamic compressors there is a dependent relationship between capacity and compression ratio, defined as the discharge pressure divided by the inlet pressure (in consistent units). Accordingly, a change in gas capacity is always accompanied by a change in the compression ratio. As the operating point of the compressor changes, the efficiency of the compression thermodynamic process will also change.

SUMMARY

In one construction, the invention provides a multi-stage compression system including a plurality of centrifugal compression stages. Each stage includes an impeller and a variable speed motor coupled to the impeller. Each variable speed motor is operable al a speed between a first speed and a second speed. The multi-stage compression system also includes a control system that is connected to each of the variable speed motors and is operable to vary the speed of each motor. The speed of each motor is varied simultaneously such that a ratio of the speed of any two variable speed motors remains constant.
In another construction, the invention provides a multi-stage compression system including a first centrifugal compressor stage having a first, impeller and a first variable speed motor. The first variable speed motor is operable at a first speed between a low speed and a high speed. The multi stage compression system also includes a second centrifugal compressor stage having a second impeller and a second variable speed motor. The second variable speed motor is operable at a second speed between a low speed and a high speed. The first speed and the second speed define a first ratio. The multi-stage compression system further includes a third centrifugal compressor stage having a third impeller and a third variable speed motor. The third variable speed motor is operable at a third speed between a low speed arid a high speed. The first speed and the third speed define a second ratio and the second speed and the third speed define a third ratio. The multi-stage compression system also includes a control system operable to synchronously vary the first speed, the second speed, and the third speed such that the first ratio, the second ratio, and the third ratio remain constant.
In yet another construction, the invention provides a method of controlling a multi-stage compression system to deliver a gas to a gas utilization system that uses the gas at a variable rate. The method includes operating a first centrifugal compressor stage al a first speed to produce a fluid flow, directing the fluid flow to a second centrifugal compressor stage, and operating the second centrifugal compressor stage at a second speed. The method also includes varying the first speed and the second speed in unison such that a ratio between the first speed and the second speed remains constant.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a cross-sectional view of a centrifugal compression stage;

FIG. 2 is a schematic of a multi-stage compression system including three centrifugal compression stages of FIG. 1; and

FIG. 3 is a performance map of the centrifugal variable speed compression stage of FIG. 1.

DETAILED DESCRIPTION

Before any embodiments of the invention are explained in detail, it is to be understood that the invention is not limited in its application to the details of construction and the arrangement of components set forth in the following description or illustrated in the following drawings. The invention is capable of other embodiments and of being practiced or of being carried out in various ways. Also, it is to be understood that the phraseology and terminology used herein are for the purpose of description and should not be regarded as limiting. The use of “including,” “comprising,” or “having” and variations thereof herein is meant to encompass the items listed thereafter and equivalents thereof as well as additional items. Unless specified or limited otherwise, the terms “mounted,” “connected,” “supported,” and “coupled” and variations thereof are used broadly and encompass direct and indirect mountings, connections, supports, and couplings. Further, “connected” and “coupled” are not restricted to physical or mechanical connections or couplings.
FIG. 1 illustrates a fluid compression module 10 (sometimes referred to as a compression stage or a compression unit) that includes a prime mover coupled to a compressor 20 and that is operable to produce a compressed fluid. In the illustrated construction, an electric motor 15 is employed as the prime mover. However, other constructions may employ other prime movers such as but not limited to internal combustion engines, diesel engines, combustion turbines, etc.
The electric motor 15 includes a rotor 25 mid a stator 30 that defines a stator bore 35. The rotor 25 is supported for rotation on a shaft 40 and is positioned substantially within the stator bore 35. The illustrated rotor 25 includes permanent magnets 45 that interact with a magnetic Held produced by the stator 30 to produce rotation of the rotor 25 and the shaft 40. In one construction, the rotor 25 is operable between about 0 RPM and 50,000 RPM, with faster speeds also being possible. Before proceeding, it should be noted that the word “between” as used herein in conjunction with discussions of operating speeds of the motors or compression stages means that the motor or compression stage is operable at any speed between the defined end points (e.g., 0 RPM, 50,000 RPM) and including the end points. Thus, a two-speed motor (i.e., one operable at 0 RPM and 50,000 RPM) is not operable between 0 RPM and 50,000 RPM. Rather, it is operable at 0 RPM and it is operable at 50,000 RPM. A motor operable between two speeds is operable at those two speeds as well as any intermediate speed between the end points. The magnetic field of the stator 30 can be varied to vary the speed of rotation of the shaft 40. Of course, other constructions may employ other types of electric motors (e.g., synchronous, induction, brushed DC motors, etc.) if desired.
The motor 15 is positioned within a housing 50 which provides both support and protection for the motor 15. A beating 55 is positioned on either end of the housing 50 and is directly or indirectly supported by the housing 50. The bearings 55 in turn support the shaft 40 for rotation. In the illustrated construction, magnetic bearings 55 are employed with other bearings (e.g., roller, ball, needle, etc.) also suitable for use. In the construction illustrated in FIG. 1, secondary bearings 60 are employed to provide shaft support in the event one or both of the magnetic bearings 55 fail.
In some constructions, an outer jacket 65 surrounds a portion of the housing 50 and defines cooling paths 70 therebetween. A liquid (e.g., glycol, refrigerant, etc.) or gas (e.g., air, carbon dioxide, etc.) coolant flows through the cooling paths 70 to cool the motor 15 during operation.
An electrical cabinet 75 may be positioned at one end of the housing 50 to enclose various items such as a motor controller, breakers, switches, and the like. The motor shaft 40 extends beyond the opposite end of the housing 50 to allow the shaft to be coupled to the compressor 20.
The compressor 20 includes an intake housing SO or intake ring, an impeller 85, a diffuser 90, and a volute 95. The volute 95 includes a first portion 100 and a second portion 105. The first portion 100 attaches to the housing 50 to couple the stationary portion of the compressor 20 to the stationary portion of the motor 15. The second portion 105 attaches to the first portion 100 to define an inlet, channel 110 and a collecting channel 115. The second portion 105 also defines a discharge portion 120 that includes a discharge channel 125 that is in fluid communication with the collecting channel 115 to discharge the compressed fluid from the compressor 20.
In the illustrated construction, the first portion 100 of the volute 95 includes a leg 130 that provides support for the compressor 20 and the motor 15. In other constructions, other components are used to support the compressor 20 and the motor 15 in the horizontal position. In still other constructions, one or more legs, or other means are employed to support the motor 15 and compressor 20 in a vertical orientation or any other desired orientation.
The diffuser 90 is positioned radially inward of the collecting channel 115 such that fluid flowing from the impeller 85 passes through the diffuser 90 before entering the volute 95. The diffuser 90 includes aerodynamic surfaces (e.g., blades, vanes, fins, etc.) arranged to reduce the flow velocity and increase the pressure of the fluid as it passes through, the diffuser 90.
The impeller 85 is coupled to the rotor shaft 40 such that the impeller 85 rotates with the motor rotor 25. In the illustrated construction, a rod 140 threadably engages the shaft 40 and a nut 145 threadably engages the rod 140 to fixedly attach the impeller 85 to the shaft 40. The impeller 85 extends beyond the bearing 55 that supports the motor shaft 40 and, as such is supported in a cantilever fashion. Other constructions may employ other attachment schemes to attach the impeller 85 to the shaft 40 and other support schemes to support the impeller 85. As such, the invention should not be limited to the construction illustrated in FIG. 1. Furthermore, while the illustrated construction includes a motor 15 that is directly coupled to the impeller 85, other constructions may employ a speed increaser such as a gear box to allow the motor 15 to operate at a lower speed than the impeller 85.
The impeller 85 includes a plurality of aerodynamic surfaces or blades that are arranged to define an inducer portion 155 and an exducer portion 160. The inducer portion 155 is positioned al a first end of the impeller 85 and is operable to draw fluid into the impeller 85 in a substantially axial direction. The blades accelerate the fluid and direct it toward the exducer portion 160 located near the opposite end of the impeller 85. The fluid is discharged from the exducer portion 160 in at least partially radial directions that extend 360 degrees around the impeller 85.
The intake housing 80, sometimes referred to as the intake ring, is connected to the volute 95 and includes a flow passage 165 that leads to the impeller 85. Fluid to be compressed is drawn by the impeller 85 down the flow passage 165 and into the inducer portion 155 of the impeller 85. The flow passage 165 includes an impeller interface portion 170 that is positioned near the blades of the impeller 85 to reduce leakage of fluid over the top of the blades. Thus, the impeller 85 and the intake housing 80 cooperate to define a plurality of substantially closed flow passages.
In the illustrated construction, the intake housing 80 also includes a flange 180 that facilitates the attachment of a pipe or other flow conducting or holding component. For example, a filter assembly can be connected to the flange 180 and employed to filter the fluid to be compressed before it is directed to the impeller 85. A pipe may lead from the filter assembly to the flange 180 to substantially seal the system after the filter and inhibit the entry of unwanted fluids or contaminates. In other embodiments, a pipe leads from the outlet of one stage to the inlet of a second stage. The pipe connects at the flange 180.
FIG. 2 illustrates three compression stages 10 a, 10 b, 10 c arranged in series to define a multi-stage compression system 205. In the illustrated construction, each compression stage 10 a, 10 b, 10 c is similar to the compression module 10 of FIG. 1. However, other constructions may use different compression modules altogether, or may include a combination of different compression module types.
For purposes of description, FIG. 2 will be described using air as the fluid being compressed. Of course one of ordinary skill in the art will realize that many other fluids can be compressed using the present system. The first stage 10 a draws in a flow of air 210 in an uncompressed state and discharges a flow of partially-compressed air 215.
The partially compressed air flow 215 passes to an inter-stage heat exchanger or cooler 220 that cools the partially compressed air flow 215 to improve the overall compression system efficiency. In the illustrated, construction, a cooling fluid 225 (e.g., cool air, water, glycol, refrigerant, etc.) flows through the heat exchanger 220 to cool the air 215.
A cooled partially compressed air flow 230 passes into the inlet, of the second stage 10 b of the multi-stage compression system 205. The second stage 10 b further compresses the air and discharges a second flow of partially compressed air 235.
The second flow of partially compressed air 235 flows through a second inter-stage heat exchanger 240 where the air is again cooled by a coolant 245 that flows through the heat exchanger 240. After passing through the second inter-stage heat exchanger 240, a cooled partially compressed air flow 250 proceeds to the third compression stage 10 c.
The third stage 10 c receives the partially compressed air flow 250 at the inlet and is operable to further compress the air to the final desired output pressure. A compressed air flow 255 is discharged from the third stage 10 c at the desired output pressure.
A final inter-stage cooler 260 may be employed following the final compression stage 10 c to cool the air before being directed to additional systems (e.g., valves, filters, dryers, etc.) or to a compressed gas utilization system 265. As with the other heat exchangers 220, 240 a flow of coolant 270 is used to cool the air before the air is discharged as a final flow of compressed air 275.
The heat exchangers 220, 240, 260 may be, for example, counter-flow heat exchangers, cross-flow heat exchangers, or parallel flow heat exchangers. In some constructions, the heat exchangers 220, 240, 260 may be plate-fin heat exchangers, shell and tube heat exchangers, or any other type operable to exchange heat between the compressed air flow 210, 235,255 and the cooling fluids 225, 245, 270, respectively. In addition, one or more of the heat exchangers 220, 240, 260 may be configured as a regenerative heat exchanger, using a compressed air flow as the cooling fluid.
The multi-stage compression system 205 also includes a check valve 280 and a blow-off valve 285, both located downstream of the last compression stage 10 c. The check valve 280 isolates the compression system 205 from the compressed gas utilization system 265 during periods when the compression system 205 is not running or when the exit pressure al the third stage 10 c is lower than the operating pressure of the utilization system 265. The blow-off valve 285 allows a portion of compressed gas to be discharged from the compression system 205 when the compressed gas generated exceeds the demand of the utilization system 265.
While FIG. 2 illustrates a three-stage compression system 205 employing a single compressor 20 at each stage 10 a, 10 b, 10 c, other systems may employ fewer or more than three stages to meet the requirements of the gas utilization system 265. In addition, some arrangements may include multiple compressors at one or more of the stages (i.e., parallel operation) to increase the capacity of the system 205. As such, the invention should not be limited to a three-stage compression system that employs only a single compressor at each stage with the stages arranged in series.
As illustrated in FIG. 2, the multi-stage compression system 205 includes three stages 10 a, 10 b, 10 c that are mechanically de-coupled from one another with each driven by a high-speed, variable speed electric motor 15 a, 15 b, 15 c. As mentioned above, the compression stages 10 a, 10 b, 10 c are arranged in series such that air passes sequentially through each stage 10 a, 10 b, 10 c before exiting the last compression stage 10 c and flowing to the compressed gas utilization system 265.
The control system 290 varies the relational speeds of the stages 10 a, 10 b, 10 c simultaneously in response to a variety of inputs. For example, in one construction, the control system 290 monitors the volumetric flow rate exiting the final stage 10 c as well as the pressure of the flow exiting the final stage 10 c and uses this data to control the speed of the individual motors 15 a, 15 b, 15 c. In other constructions, the control system 290 monitors the pressure at the gas utilization system 265 and/or the rate of use at (he gas utilization system 265.
The control system 290 may include sensors for sensing pressure, temperature, and flow rate al various points within the system 205. Additionally, the control system 290 may include control logic algorithms and associated hardware to run the algorithms and monitor the system parameters. Furthermore, the control system 290 may include supplemental hardware for powering the electric motors 15 a, 15 b, 15 c and regulating their speeds m response to the measured system parameters. The control system 290 is designed to regulate the performance of the centrifugal compression stages 10 a, 10 b, 10 c without the use of an inlet throttling valve. However, in some constructions, a throttling valve may still be provided at the inlet of the first compression stage 10 a. In one control system 290 sensors are employed to measure the power (kW) consumed by one or more of the motors. This power consumption is then used to control the speed of the motors.
While each stage 10 a, 10 b, 10 c is mechanically decoupled from the other stages, preferred constructions employ a control system 290 that varies the speed of the motors 15 a, 15 b, 15 c synchronously. For example, in one construction, the first motor 15 a operates at a first speed between a low speed and a high speed, the second motor 15 b operates at a second speed between a low speed and a high speed, and the third motor 15 c operates at a third speed between a low speed and a high speed. In this example, the speed of each motor 15 a, 15 b, 15 c is varied by the control system 290 at the same time such that the ratio of the speeds of any two motors 15 a, 15 b, 15 c remains constant. Thus, if the speed of the first motor 15 a is doubled, the speeds of the second motor 15 b and the third motor 15 c are also doubled. In one construction, the first speed, the second speed, and the third speed are substantially the same. However, different speeds could be employed if desired.
During the compression process, each stage 10 a, 10 b, 10 c imparts energy to the air in the form of pressure and motion. While the mass flow through the compression system 205 remains practically constant, with the exception of some amount of air that escapes through various seals, the volume of the air is reduced from stage to stage as the pressure increases. Customarily, the gas leakage flow is limited to between 1 percent and 2 percent of the volumetric inlet capacity of the compressor 20.
A typical centrifugal compressor performance map 300 is shown in FIG. 3. FIG. 3 can be interpreted as the characteristic performance map of a multi-stage centrifugal compression system, or as the characteristic performance map of a single stage of compression, although the specific values for each of the curves will be different depending on which map it is. From FIG. 3, it can be readily deduced that if the process demand departs from an operating point O_o, a variation in the operating speed of the multi-stage compression system 205 may satisfy the new process demand while maintaining a high level of operating efficiency.
For a given fixed rotational speed (shown in FIG. 3 as constant speed lines 305) of a centrifugal compressor stage, a relationship exists between capacity and stage pressure ratio. FIG. 3 suggests that an increase in capacity is accompanied by a decrease in pressure ratio. Conversely, a decrease in capacity corresponds to an increase in the pressure ratio of the compression stage. When multiple stages are connected in series the pressure of the gas is incrementally increased. If the inlet of an intermediate stage coincides with the discharge of the immediately upstream stage, and the discharge of any stage coincides with the inlet of the following stage, the pressure ratio of each stage accounts for interstage losses. The pressure ratio of a complete multi-stage compressor can be expressed as the product of the pressure ratios, R_i, of the individual stages. The discharge pressure from the last stage of compression, P_disch, would then be equal to the product of the compressor pressure ratio, R_comp, times the inlet pressure, P_inlet, to the compressor, as indicated by the following equation:
P _disch=(R ₁ ×R ₂ ×, . . . , ×R _N)×P _inlet =R _comp ×P _inlet (1)
The demand of the downstream compressed gas utilization system 265 dictates the How and pressure conditions at which the compression system 205 operates. An increase in system pressure indicates a supply of compressed gas that exceeds the demand of the receiving system (i.e., compressed gas is being delivered taster than it is being used). Similarly, a decrease in receiving system pressure triggers a request of a greater capacity from the compression system 205 to maintain, in the case of a constant pressure process, the target operating pressure (i.e., gas is being used faster than it is being delivered). When Equation (1) is applied to a constant speed compressor, with performance corresponding to one speed line 305 in the diagram of FIG. 3, and the condition of a constant discharge pressure is imposed, i.e. P_disch=Constant, then a required capacity change to accommodate the process demand will affect the pressure ratio of the compression system 205 since the pressure ratio of each stage 10 a, 10 b, 10 c is uniquely linked to a given capacity, it follows that since the term on the left-hand side of Equation (1) is constant, and the pressure ratio of the compression system 205 physically changes, then the inlet pressure to the first compression stage 10 a assumes a value so that the product in Equation (1) remains, in fact, constant.
As an example, if the capacity handled by the compression system 205 is reduced in order to counteract a contingent increase in system pressure, the pressure ratio of the compression system 205 increases because, based on the performance Characteristics of centrifugal compressors, a decrease in capacity corresponds lo an increase in compressor pressure ratio. Accordingly, to maintain constant pressure at the discharge of the compression system 205, the inlet pressure to the compression system 205, P_inlet ^(new), assumes a new value equal to:
$\begin{matrix} \frac{P_{disch}}{R_{comp}^{(New)}} = P_{inlet}^{(New)} & (2) \end{matrix}$
where:
R_comp ^(New)=Compressor pressure ratio corresponding to the required flow
P_inlet ^(New)=Updated, new inlet pressure at the first stage of compression to maintain the required constant discharge pressure
The change in the inlet pressure of a first stage of compression, in the case of an open to atmosphere compression system, is typically obtained in prior art systems by interposing a throttling valve between the atmospheric environment and the inlet of the first stage of compression. Following the example at hand, a calibrated closure of the inlet throttling valve introduces a pressure drop, thus reducing the pressure at the inlet of the first compression stage. The new reduced inlet pressure that satisfies Equation (2), when multiplied by the increased pressure ratio required to achieve a smaller compressor capacity, yields the required constant pressure at the discharge of the compression system. This is the most common regulation system for fixed speed compressors, and while it properly controls the capacity of a centrifugal compressor, it also introduces an energy loss because of the presence of the inlet throttling valve.
However, an unaltered pressure ratio of the compression stages 10 a, 10 b, 10 c, under a changing capacity requirement, can also be satisfied by modifying the operating speed of the stages 10 a, 10 b, 10 c. Accordingly, the overall pressure ratio of the compression system 205 can be maintained as the capacity is adjusted to the demand of the utilization system 265. Since the discharge pressure of the compression system 205 is maintained, under changing capacity requirements, on the basis of the relationship that is uniquely related to the pressure ratio, the capacity, and the rotational speed of the compression stages 10 a, 10 b, 10 c, there is no requirement to alter the inlet pressure to the first stage 10 a with a regulating inlet throttling valve. In fact, with reference to Equation (2), if the compressor pressure ratio is maintained constant, then the inlet pressure to the first stage 10 a does not need to be changed to maintain a value for the discharge pressure. Because of the change in capacity, the inlet pressure of the compression system 205 may actually change because of changes in the pressure losses that are linked to the contingent capacity. Nonetheless, in this context, it is emphasized that the control system 290 does not require the first stage inlet pressure to be varied by means of a throttling valve, and accordingly there is no throttling valve. It should be noted that the removal of the inlet throttling valve directly improves the efficiency of the first stage 10 a, reduces control software complexity, eliminates hardware, and improves costs as compared to fixed speed systems that include the throttling valve.
Each stage 10 a, 10 b, 10 c in the multi-stage compression system 205 is designed to handle fluid having a certain inlet pressure range, but particularly a defined intake volume rate range. Because the stages 10 a, 10 b. 10 c are in series, each stage 10 a, 10 b, 10 c is flow-matched to the others so as to allow for the proper transfer of fluid from one stage to the next. Accordingly, if for some reason one stage is not in the condition to accept the flow delivered by an upstream stage, or does not deliver the flow at the intended pressure and volume to the next stage, a disruption in the operation of the multi-stage compression system 205 may occur, thereby resulting in an upset in the performance of all the stages 10 a, 10 b, 10 c in the compression system 205. Because of the gas dynamics accompanying the behavior of centrifugal compressors, a condition of flow instability may quickly develop which requires the unloading of the compression system 205 to properly handle and suppress the unstable situation.
The performance of the compression system 205 is particularly sensitive to the variation in rotational speed of the stages 10 a, 10 b, 10 c. Compressor stage capacity varies linearly with speed, head changes vary proportionally to the square of the speed, and power varies as the cube of the stage speed. Because of the aerodynamic coupling between compressor stages 10 a, 10 b, 10 c in series, the approach of independently changing the speed of each stage 10 a, 10 b, 10 c, when the stages 10 a, 10 b, 10 c are mechanically de-coupled, requires a significant level of sophistication and control logic complexity, The operating conditions of each compression stage 10 a, 10 b, 10 c is monitored in terms of gas pressure, temperature, and flow, so as to predict the speed change for one stage in conjunction with the speed of the other stages, while not upsetting the internal operation of each compression stage 10 a, 10 b, 10 c and while satisfying the demand of the utilization system 265.
The rate of change in fluid flow demand of the utilization system 265 adds complications to an approach of independently regulating the speeds of the stages 10 a, 10 b, 10 c because background calculations are performed before attempting the adaptation of the compression system 205 to the variation in the operating conditions of the utilization system 265. A significant time delay may be introduced between the acknowledgement that the process conditions are changing and an action is taken to satisfy the process demand. Moreover, a predetermined databank, must be available to the computational control algorithms so as to reasonably predict the appropriate speed regulation of the independently driven centrifugal compressor stages 10 a, 10 b, 10 c present in the compression system 205.
An explorative type approach may be employed, where changes are made in a semi-static compressor operating condition so as to prevent transient driven upsets within the compressor system while continuing to satisfy the utilization system demand in terms of discharge pressure and flow rate. Such an approach may not satisfy required rapid changes in the operating conditions of the compression system 205 to respond to the demand of the compressed gas utilization system 265.
The constructions discussed herein approach the speed regulation of the compression system 205 by adjusting the contingent operating speed of each stage 10 a, 10 b, 10 c synchronously, while maintaining a predetermined constant ratio between the speeds of each stage 10 a, 10 b, 10 c. This means that the speed of each stage 10 a, 10 b, 10 c is modified at the same time and the ratio between the rotational speeds is kept constant al all times. In one construction, a predetermined value of the speed ratio is equal to one. In this case, a speed ratio of one corresponds to the condition where two referenced stages operate at the same rotational speed. In other embodiments, the ratios are equal to values oilier than one, meaning the two referenced stages are operating at different rotational speeds.
During the act of regulation, as the stage speeds are synchronously changed, the ratio between the speeds will remain constant as predetermined by the individual characteristics of the compressor stages. Nonetheless, in the control algorithms, the speed ratios can be defined as variables that can be changed from one compressor configuration to another. For example, one set of simultaneous constraint equations that characterize the control approach can be summarized symbolically as follows: Equations (3a) impose the condition that the speed ratio, and consequently the rotational speed change of one stage with respect to all the other stages, is constant during the compressor control action.
$\begin{matrix} \frac{Δ N_{1}}{Δ N_{2}} = K_{1 - 2} \frac{Δ N_{1}}{Δ N_{3}} = K_{1 - 3} ⋮ \frac{Δ N_{1}}{Δ N_{n}} = K_{1 - n} ⋮ \frac{Δ N_{n - 1}}{Δ N_{n}} = K_{(n - 1) - n} & (3 a) \\ Δ N_{1} = Δ N_{1} δ Δ N_{2} = Δ N_{2} δ ⋮ Δ N_{n} = Δ N_{n} δ & (3 b) \\ δ = 0, 1 & (3 c) \\ Abs (P_{disch} - P_{demand}) \leq ɛ_{P} (Constant Pressure Control) & (3 d) \\ Abs (Q_{disch} - Q_{demand}) \leq ɛ_{Q} (Constant Delivered Capacity Control) & (3 e) \end{matrix}$
where:
ΔN_i=Speed change of the ith stage
K_i−j=Speed ratio of the ith stage with respect to the jth stage. The speed ratio is a real number in a mathematical sense,
δ=a Kronecker delta function which can be interpreted as a control decision variable suggesting that stage speed variations must occur synchronously.
ε_P=Error related to pressure
ε_Q=Error related to capacity

P=Pressure

Q=Capacity

Equations (3b) identify the synchronous occurrence of the rotational speed change with variable δ assuming a value of zero when no speed regulation is in process, and assuming a value of one when compressor control, with reference to stage speed changes, is required. Equations (3d) and (3e) impose the condition that the speed change satisfies the demand of the compressed gas utilization system 265 within a certain tolerance.
The compressor control scheme is enhanced by an aerodynamic design of the compression stages 10 a, 10 b, 10 c that accommodates a synchronous variation of the rotational speed of the stages 10 a, 10 b, 10 c while insuring stable changes in compressor head, pressure ratio, and capacity. Such aerodynamic design elements include evenly distributed impeller blade loading, advanced 3-D impeller blade shapes, and a 3-D diffuser design. Thus, the illustrated construction operates in a way that avoids boundary conditions at the inlet or at the discharge of a stage that may cause the surge limit 310 or choke limit 315, shown in FIG. 3, to be active constraints.
FIG. 3 also shows paths 320 along which the efficiency of the compression process is constant. The highest efficiency of the stages 10 a, 10 b, 10 c is planned at the nominal operating conditions and, in the figure, is identifiable by the region at the center of the constant efficiency paths 320. The deterioration in efficiency can be kept to a minimum as the operating point O_odeparts from the design point, such that, for dynamically similar stages, a synchronous stage speed change, while maintaining a constant ratio between the speed of each stage, does not significantly deteriorate the efficiency of the multi-stage compression system 205.
Also, the design of dynamically similar stages, in aerodynamic terms, allows stable operation during speed variations, without incurring transient instability issues. The aerodynamic design of the stages 10 a, 10 b, 10 c augments the performance of the control system 290 by resulting in a modest deterioration of the design optimal efficiency within a significant speed range.
Further, a control inlet throttling valve is not required in the multi-stage compression system 205, thus reducing the cost and complexity of the system 205.
Finally, compressor unload can be achieved by synchronously reducing the speeds of the stages 10 a, 10 b, 10 c while maintaining a constant relative speed ratio. Opening the blow-off valve 285 accommodates the transient unload event of the compression system 205, as the check valve 280 isolates the compression system 205 from the utilization system 265. The control system 290 may also allow the complete shut-down of the compression system 205 in the event that the utilization system 265 does not require the operation of the compression system 205.
Various features and advantages of the invention are set forth in the following claims.

Claims

What is claimed is:

1. A multi-stage compression system comprising:

a plurality of centrifugal compression stages, each stage including an impeller and a variable speed motor coupled to the impeller, each variable speed motor operable at a speed between a first speed and a second speed; and

a control system connected to each of the variable speed motors and operable to vary the speed of each motor, the speed of each motor being varied simultaneously such that a ratio of the speed of any two variable speed motors remains constant.

2. The multi-stage compression system of claim 1, wherein a fluid flow passes through the plurality of centrifugal compression stages sequentially.

3. The multi-stage compression system of claim 1, wherein the plurality of centrifugal compression stages are mechanically de-coupled from each other.

4. The multi-stage compression system of claim 1, further comprising at least one heat exchanger, the at least one heat exchanger positioned to receive a fluid flow from at least one of the plurality of centrifugal compression stages.

5. The multi-stage compression system of claim 1, wherein the control system includes at least one sensor positional to measure a parameter of a fluid flow passing through the plurality of centrifugal compression stages.

6. The multi-stage compression system of claim 5, further comprising a valve disposed between the plurality of centrifugal compression stages and a gas utilization system, the valve operable to discharge a portion of the fluid flow from the multi-stage compression system.

7. The multi-stage compression system of claim 6, wherein the control system varies the speed of the variable speed motor and operates the valve at least partially in response to the measured parameter.

8. The multi-stage compression system of claim 5, wherein the parameter is at least one of a pressure, a temperature, and a mass flow rate.

9. The multi-stage compression system of claim 1, wherein a mass How rate of the plurality of centrifugal compression stages is substantially constant.

10. The multi-stage compression system of claim 1, wherein a mass flow rate of the plurality of centrifugal compression stages is variable in response to use of a fluid flow at a gas utilization system.

11. The multi-stage compression system of claim 1, wherein a first variable speed motor operates at a different speed than a second of the variable speed motors.

12. A multi-stage compression system comprising:

a first centrifugal compressor stage including a first impeller and a first variable speed motor, the first variable speed motor operable at a first speed between a low speed and a high speed;

a second centrifugal compressor stage including a second impeller and a second variable speed motor, the second variable speed motor operable at a second speed between a low speed and a high speed, the first speed and the second speed defining a first ratio;

a third centrifugal compressor stage including a third impeller and a third variable speed motor, the third variable speed motor operable at a third speed between a low speed and a high speed, the first speed and the third speed defining a second ratio and the second speed and the third speed defining a third ratio; and

a control system operable to synchronously vary the first speed, the second speed, and the third speed such that the first ratio, the second ratio, and the third ratio remain constant.

13. The multi-stage compression system of claim 12, wherein the first centrifugal compressor stage, the second centrifugal compressor stage, and the third centrifugal compressor stage are arranged in series.

14. The multi-stage compression system of claim 12, wherein the centrifugal compressor stages are mechanically de-coupled from each other.

15. The multi-stage compression system of claim 12, further comprising a first heat exchanger positioned between the first centrifugal compression stage and the second centrifugal compression stage, a second heat exchanger positioned between the second centrifugal compression stage and the third centrifugal compression stage, and a third heat exchanger positioned between the third centrifugal compression stage and an air utilization system.

16. The multi-stage compression system of claim 12, wherein the control system includes at least one sensor positioned to measure a parameter of a fluid flow passing through the centrifugal compression stages.

17. The multi-stage compression system of claim 16, further comprising a valve disposal between the third centrifugal compression stage and an air utilization system, the valve operable to discharge a portion of the fluid flow from the multi-stage compression system.

18. The multi-stage compression system of claim 17, wherein the control system varies the speed of the first variable speed motor, the second variable speed motor, and the third variable speed motor and operates the valve in response to the measured parameter.

19. The multistage compression system of claim 16, wherein the parameter is at least one of a temperature, a pressure, and a mass flow rate.

20. The multistage compression system of claim 12, wherein a mass flow rate of the centrifugal compression stages is substantially constant.