TECHNICAL FIELD OF THE INVENTION

This invention relates to centrifugal compressor systems, and more particularly to centrifugal compressor systems for high pressure and high-pressure ratio applications .

BENEFIT OF PROVSIONAL FILING DATE

This patent application claims the benefit of the filing date, March 29, 2019, of U.S. Provisional Patent Serial No. 62/826167, entitled "Centrifugal Compressor with Piston Intensifier" .

BACKGROUND OF THE INVENTION

A Hydrogen Refueling Station (HRS) is an

infrastructure designed for filling a vehicle with hydrogen fuel. The HRS can be part of a production unit where hydrogen is produced or can be remote from

production. If the H2 is produced on site or is delivered to the station at an intermediary pressure or in liquid state, the HRS requires an intermediary storage and a compression system.

The compressor must be capable of bringing the hydrogen to a desired gas pressure level. Hydrogen compression is used to overcome the pressure difference between storage (from 50 to 200 bar) and refueling (up to 1,000 bar) . The refueling process should not exceed short target time of as little as three to five minutes.

Because the vehicle's fuel cell is operated with pure hydrogen, it is important that no contamination with lubricants occurs during compression.

Currently available compressors fail to meet the order-of-magnitude higher dispensing needs of heavy-duty (HD) applications and are not designed to scale reliably or cost-effectively. Hydrogen Refueling Station (HRS) compressors have fundamentally different needs and requirements than traditional hydrogen applications, and therefore compression systems should be designed to be fit for purpose.

Existing compression technologies include

reciprocating and hydraulic piston compressors that suffer from low reliability and extremely short

maintenance intervals due to wear of oil-free components and oil isolation seals required to meet high hydrogen purity requirements. Diaphragm compressors are hermetically sealed but have demonstrated notable

diaphragm degradation due to hydrogen chemical attacks and physical damage arising from rapidly changing air bubble solubility of the hydraulic fluid caused by on-off cycling. In addition to reliability concerns, a major challenge with conventional technologies is the ability to scale to the high throughput requirements for HD refueling with low suction pressures, which would require massive pistons or diaphragm heads. The scaling of existing technologies is not a trivial or even

necessarily logical approach to meet the needs of the HD market .

Centrifugal compressors are typically a reliable technology of choice for high-flow high-pressure (HP) process applications, but the HP ratio of a refueling compressor for an integrated production-storage-refueling system presents significant challenges. Due to hydrogen's very low molecular weight, typical compressor impeller tip speeds up to 400 m/s would require nearly 50 stages of compression. Low volume flow rates at the final stages would result in stage designs with low efficiency and manufacturing challenges due to small flow passage sizes.

BRIEF DESCRIPTION OF THE DRAWINGS

A more complete understanding of the present

embodiments and advantages thereof may be acquired by referring to the following description taken in

conjunction with the accompanying drawings, in which like reference numbers indicate like features, and wherein:

FIG. 1 illustrates one stage of the compressor system.

FIG. 2 illustrates example specifications for a compressor system having 6 - 8 stages.

SUMMARY

The following description is of gas compression system that comprises a centrifugal compressor in fluid connection in series with a piston intensifier. Fluid flow from the centrifugal compressor is provided to the piston intensifier. The centrifugal compressor drives the piston intensifier to raise the pressure of the fluid, particularly using at least a portion of the fluid from the centrifugal compressor as the working fluid for the piston intensifier. At least a portion of the

remaining portion of fluid from the centrifugal

compressor is provided to the intensifier compressor as the fluid being compressed to the desired pressure. That is, the fluid flow output from the centrifugal compressor is divided into at least two inputs being provided to the intensifier compressor: (1) as working fluid and (2) as the fluid being compressed.

The description is also of methods for compressing a fluid using a fluid compression system comprising

centrifugal compression and piston intensifier. One exemplary method includes providing a fluid to a

centrifugal compressor, operating the centrifugal compressor to compress the fluid and provide a first compressed fluid as an output fluid flow, providing at least a first portion of the output fluid flow of the centrifugal compressor as a working fluid to an

intensifier to drive at least a portion of the

compression operation by the intensifier, providing a second portion of the output fluid flow of the

centrifugal compressor as a production fluid to the intensifier, said production fluid to be compressed by the compression operation of the intensifier driven at least by the working fluid from the output fluid flow of the centrifugal compressor, providing at least a portion of the compressed production fluid as an output fluid flow from the intensifier at a pressure higher than the pressure of the first compressed fluid, and recycling at least a portion of the working fluid back to the

centrifugal compressor for continued use as the working fluid .

DETAILED DESCRIPTION OF THE INVENTION

The following description is directed to a

compressor system for high pressure and high-pressure ratio gas transfer applications. Although this

description is written in terms of hydrogen refueling applications, the compressor system may also be used with other gasses, such as helium, natural gas, air, nitrogen, carbon dioxide. Many other applications are possible for transferring gas from a source at one pressure to a destination at a higher pressure. The compression system is compact and scalable for various applications.

One application is for high throughput hydrogen fueling for medium and heavy-duty transportation. The system has application at both refueling stations and production sites that distribute to refueling sites. The compression system may be used to fill vehicles (on board storage) or to fill intermediate storage.

The compression system combines a centrifugal compressor with a two-stage hermetically-sealed piston intensifier for reliably delivering 400+ kg/hr of

hydrogen at 530 bar. The system may be described as a high-pressure ratio and low flow system, with a pressure ratio of 53:1 and flow of 400 kg/hr being described for purposes of example.

The piston intensifier enables a high system

pressure ratio by using excess hydrogen flow from the centrifugal compressor as a motive fluid, thus

eliminating concerns of leakage or contamination from liquid or pneumatic piston concepts. This solution also offers improved reliability over existing compressor solutions by reducing the dynamic stresses associated with high-speed reciprocating equipment.

FIG. 1 illustrates a hybrid compression system in accordance with the invention. Only one stage is shown in FIG. 1; various applications of the system may call for multiple stages. A typical implementation for hydrogen fueling might be two double-acting pistons for every compressor. Depending on the application, other

configurations are possible.

FIG. 2 (and the Table below) set out examples of sizing and performance specifications for a 6 - 8 stage system with two pistons per compressor stage. As stated above, the number of stages and number of pistons per compressor stage may vary depending on the application.

The compressor system has a low-pressure high-flow centrifugal compressor that simultaneously provides both pre-compression of the supply gas and actuation gas for a high-pressure piston intensifier. This configuration is advantageous over traditional reciprocating or diaphragm compressor solutions as well as conventional centrifugal- only compressor solutions in that it (a) minimizes leakage to atmosphere via low-leakage dry gas seals on the centrifugal compressor and a hermetically-sealed intensifier, (b) eliminates the risk of contamination in the intensifier via hydrogen motive gas, (c) maximizes the system reliability by relaxing piston seal leakage requirements and leveraging proven centrifugal compressor technology including recent advancements for hydrogen service, (d) minimizes compression power through

intercooling between compressor and intensifier stages, and (e) enables near-isentropic efficiency in the

intensifier cylinders via slow piston speeds and minimal clearance volumes.

In the example of FIG. 1, the compressor system has an inlet pressure of 10 bar and delivers hydrogen to the piston intensifier at 40 bar. This inlet pressure is typical of today's hydrogen production systems, also some production may discharge at slightly lower pressures.

The compressor suction and discharge lines are connected to the piston intensifier, which uses the 4:1 pressure ratio and approximately 90% of the mass flow produced by the compressor to drive a two-stage double acting high-pressure piston (only one stage is explicitly illustrated) .

The piston intensifier delivers the remaining 10% of the flow (400 kg/hr) to 530 bar storage. The 530-bar discharge pressure is for purposes of example - actual target discharge pressures may vary depending on the application .

At least one shaft end seal on the drive end is necessary as the required motor power and compressor speed exceed the existing capacity of high-speed motors. However, the design will potentially incorporate gas film bearings to improve compactness, eliminate a lubrication oil system, and eliminate the non-drive-end shaft seal. Gas film bearings have been evaluated in demonstrated high load capacities at elevated pressures, with test experience for both journal and thrust gas bearings at pressures up to 70 bar.

The piston intensifier motion is controlled through external valves that alternately connect the inboard and outboard cylinders to the centrifugal compressor suction and discharge lines. Proper timing of the piston motion is managed using the valves shown in FIG. 1.

As indicated by the different valve shadings in FIG. 1, the valves alternate their open/closed states for the double acting actuation of the actuation piston. In other words, the non-shaded valves are open while the shaded valves are closed and vice versa. The supply lines to and from the piston shown as solid lines are open while the supply lines shown as dashed lines are closed, and vice versa. Flow to and from the high-pressure chamber of the piston is controlled by check valves in the high-pressure chamber (not shown) .

Referring to the example of FIG. 1, hydrogen exits a source system at 10 bar pressure. It enters via the inlet line of the centrifugal compressor at 10 bar and exits at approximately 4000kg/hr at 40 bar. Ten percent (10%) of the 40 bar hydrogen enters the piston intensifier to be compressed. The remaining 90% of flow is diverted through a working fluid loop to drive the pistons. This diverted flow operates in alternating cycles through each side of the working cavity of the piston. Actuation of control valves on each side of the working cavity effectively swaps the high-pressure and low-pressure sides of the working cavity to drive the pistons. Proper timing of the piston motion will be managed via the valves. Expended working fluid will route back to the 10 bar line of the to be recycled back into the centrifugal compressor. The 10% of H2 flow that is used for production fluid enters the high-pressure cavity to be compressed by the action of the working fluid controlled by the actuating control valves. While this system is designed for 10 bar inlet pressure and 550 bar outlet pressure, the discharge pressure can be controlled downward by opening the discharge valve on the intensifier at a lower pressure.

Although initial sizing was performed with a two- stage piston intensifier, additional stages will allow intercooling and may reduce net system power. The

operating speed and surface velocity across the

intensifier will depend upon the final sizing, but preliminary results shown in FIG. 2 indicate that a 50 mm HP piston diameter (273 mm actuation piston diameter) keeps the peak surface velocity below a conventional 1.0 m/s limit of existing rod and piston seals at the

predicted discharge temperatures below 160°C.

The centrifugal compressor design target may be implemented at commercial technology limits for gearbox speed ratios, impeller tip speeds, and shaft end seal speeds in order to minimize overall system size and cost. Detailed design, analysis, and materials selection are required to ensure an efficient and reliable system including assessment of rotor dynamics of the high stage count rotor and aerodynamics to maintain efficiency in the latter stages with low flow coefficients. The

intensifier design calls for seal selection and

optimization to minimize size via increasing operating speed (within rod seal surface speed limits) . Transient performance of the coupled centrifugal-intensifier system may be analyzed via simulations and prototype testing. Design work includes design for the full-scale system and detailed design for reduced-scale test

prototypes that target specific technology gaps

identified for the full-scale system. Design of the full- scale system includes one-dimensional aerodynamic design and rotor dynamic analysis of the compressor and pressure vessel design to develop detailed two-dimensional layouts of the compressor and piston intensifier. Component

(couplings, bearings, seals, valves, gearbox, motors, etc.) selection ensures design compatibility with

existing commercial technologies and to support full system cost models.

A dynamic model of the piston intensifier coupled with the centrifugal compressor may be used to advance and optimize the piston design and valve timing and avoid unacceptable system dynamics (e.g., compressor surge) during operation. Detailed design of the reduced-scale test prototypes will include three-dimensional flow path designs for the compressor stage, including computational fluid dynamic and finite element analyses of the critical machine components.

The stop/start sequence of the compressor system may be designed for the duty cycles expected at hydrogen refueling stations. An example is to startup the

centrifugal compressor in recycle, transition to the first stage piston in recycle, and then bring the final stage piston online. Centrifugal compressors can be designed for a very large number of starts/stops . With regard to the piston, the use of hydrogen as both

actuation and process fluid allows large seal clearances and/or hydrostatic bearings to minimize wear. The compressor system is easily scaled up to

increase throughput. The centrifugal compressor is scaled by increasing the (currently small) flow coefficients for each stage. One alternative for scale-up of the piston intensifier is to use additional piston cylinders in parallel .

A feature of the invention is the positive

displacement geometry of the piston intensifier. The piston intensifiers use hydrogen in both pistons, thus using hydrogen for both actuation and process. This allows them to use higher leakage non-contacting piston seals and/or hydrostatic bearings to minimize piston/seal wear .

For the centrifugal compressor:

Inlet pressure (bara) 10.0

Inlet temperature (degrees C) 20.0

Discharge pressure (bara) 40.0

Maximum stage discharge temperature

(degrees C) 114.6

Mass flow (kg/hr) 3469

Power (MW) 2.33 - 2.73

Impeller tip diameter (mm) 61.6

Tip speed (m/s) 700

Minimum flow coefficient 0.015

For the piston intensifier:

Stage ½ discharge pressures (bara) 145/530 Stage ½ discharge temperature (degrees C) 158 Discharge mass flow (kg/hr) 400

Actuation mass flow (kg/hr) 3069

Stage ½ HP piston diameter (mm) 300/100 Stage ½ Acutation piston diameter (mm) 636/209 Stroke (mm) 500 100

Speed (cycles per minute) 18 35

Maximum piston velocity (m/s) 0.950