Back Pressure Turbine

There are four back pressure turbines for generation of electricity from VHP to HP and one back pressure turbine between HP and LP steam levels.

From: Computer Aided Chemical Engineering , 2011

Advanced ultra-supercritical pressure (A-USC) steam turbines and their combination with carbon capture and storage systems (CCS)

H. Nomoto , in Advances in Steam Turbines for Modern Power Plants, 2017

21.4.2.5 Using a back-pressure turbine between the exhaust of the intermediate-pressure turbine and the inlet of the low-pressure turbine

Placing a back-pressure turbine between the intermediate turbine and the LP turbine, which is illustrated as the turbine BPT in Fig. 21.12, makes the efficiency penalty smaller because inefficient throttling can be avoided. The expansion line by the turbine BPT is shown from PBin to PBout and as Case (C) in Fig. 21.13 (B), in which the back-pressure turbine recovers output without sacrificing heat drop. Though the efficiency sacrifice is smaller compared with the case that throttles the inlet pressure of the LP turbine, capital cost will be higher because another turbine is needed.

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Electricity generation

I G Crow BEng, PhD, CEng, FIMechE, FIMarE, MemASME , K Shippen BSc, MBA, CEng, MIMechE , in Plant Engineer's Reference Book (Second Edition), 2002

22.3.1 Steam turbines for CHP

Many industrial processes require electrical power and heat. This heat is often provided from large quantities of low-pressure steam. In this section it is demonstrated that a thermal power station gives up very large quantities of heat to the cooling water in the condenser. For this purpose, the steam pressure in the condenser is usually at the lowest practical pressure (around 0.05 bara) to achieve maximum work output from the turbine.

However, if the turbine back pressure is raised to above atmospheric pressure so that the turbine exhaust steam can be transported to the process heat load then the steam will give up its latent heat usefully rather than reject this to the condenser cooling water. Although the steam turbine output is reduced, the overall efficiency is increased significantly as the generated steam is used to provide both heat and electrical power.

The alternative (but equally appropriate) logic is that a factory may use steam from low-pressure boilers. By increasing the steam pressure and then expanding this through a steam turbine to the desired process pressure additional electrical power can be provided.

An example is shown in Figure 22.14. By raising steam at high pressure (say, 60 bara and 540°C) and then expanding this through a turbine to the process steam pressure requirements of 3 bara then useful work can be done by the turbine for generation of electrical power. For this example each kg/s of steam gives 590 kW of electrical power.

Figure 22.14. The back-pressure steam turbine in CHP application

There are several types of steam turbine that can be used to meet widely varying steam and power demands. They can be employed individually or in combination with each other.

22.3.1.1 Back-pressure turbine

The simple back-pressure turbine provides maximum economy with the simplest installation. An ideal backpressure turbogenerator set relies on the process steam requirements to match the power demand. However, this ideal is seldom realised in practice. In most installations the power and heat demands will fluctuate widely, with a fall in electrical demand when steam flow, for instance, rises.

These operating problems must be overcome by selecting the correct system. Figure 22.15 shows an arrangement which balances the process steam and electrical demands by running the turbo-alternator in parallel with the electrical supply utility. The turbine inlet control valve maintains a constant steam pressure on the turbine exhaust, irrespective of the fluctuation in process steam demand.

Figure 22.15. A back-pressure turbo-alternator operating in parallel with the grid supply

This process steam flow will dictate output generated by the turbo-alternator and excess or deficiency is made up by export or import to the supply utility, as appropriate. The alternative to the system in Figure 22.15 is to use a back-pressure turbine with bypass reducing valve and dump condenser, as shown in Figure 22.16.

Figure 22.16. A back-pressure turbine with PRDs valve and dump condenser

On this system the turbine is speed controlled and passes steam, depending on the electrical demand. The bypass reducing valve with integral desuperheater makes up any deficiency in the steam requirements and creates an exhaust steam pressure control. Alternatively, any surplus steam can be bypassed to a dump condenser, either water or air cooled, and returned to the boiler as clear condensate.

22.3.1.2 Pass-out condensing turbines

If the process steam demand is small when compared with the electrical demand then a pass-out condensing turbine may provide the optimum solution. Figure 22.17 illustrates a typical scheme, which consists of a backpressure turbine. This gives operational flexibility of the back-pressure turbine with improved power output.

Figure 22.17. The pass-out condensing turbine

22.3.1.3 Back pressure with double pass-out

Many industries require process steam at more than one pressure, and this can be done by use of a backpressure turbine supplying two process pressures (see Figure 22.18).

Figure 22.18. The double pass-out turbo-alternator

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Drivers

Grant Musgrove , ... Matt Taher , in Compression Machinery for Oil and Gas, 2019

Selection

The selection process for a mechanical drive steam turbine depends on the type of process, the available steam conditions, the required power, and the speed range. In addition, there may be steam rate and possibly casing connection size requirements.

Basic steps for ideal back pressure turbine:

1.

Establish a TSR (from inlet, exhaust steam conditions, and required power).

2.

Determine isentropic heat drop across turbine.

3.

Divide the total isentropic heat drop by the optimum heat drop per stage (typically available from OEM).

4.

This value represents the number of stages needed to make power at isentropic conditions.

5.

Assume efficiency and multiply by the total isentropic heat drop to obtain the actual heat drop.

6.

Subtract actual heat drop from isentropic heat drop to obtain exhaust enthalpy.

Other checks needed for a selection include:

•

Blade mechanical checks—airfoil profile/nozzle selection, blade material selection, Goodman/Campbell review, root design selection, momentary speed limit check.

•

Moisture erosion—erosion protection for rotating and stationary components.

•

Casing design—HP and LP casing designs based on the pressure and temperature limitations.

•

Casing connections—velocity limits and pressure rating.

•

Shaft end size—check to determine torque capability of shaft end.

•

Rotor design—lateral/torsional critical speed analysis.

•

Journal bearing—journal bearing load review, bearing metal temperature prediction.

•

Thrust check—thrust prediction across all operating points maximum thrust bearing limitations.

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Steam turbine cycles and cycle design optimization

A. Ohji , M. Haraguchi , in Advances in Steam Turbines for Modern Power Plants, 2017

2.3.2 Back pressure turbine

The back pressure turbine is used for supplying process steam to the facilities in private-use power producers. This type of steam turbine supplies not only electricity but also the process steam to the facilities. In other words, exhaust steam pressure is set to be the demanded pressure from the facility needs or outside needs. In the back pressure turbine, an effective heat drop will be small as shown in Fig. 2.13, therefore, the turbine output will be also small.

Figure 2.13. Heat drop of back pressure turbine.

In the case where large amounts of steam are required by facilities for process steam, high thermal efficiency will be expected, which means the back pressure turbine will give advantage to private power utilities. And as the back pressure turbine consists of fewer turbine stages with simple structure and small exhaust parts, this results in lower equipment costs.

The back pressure turbine (or the extraction back pressure turbine) is adopted in many facilities such as oil refineries, petrochemical, paper-pulp, fiber, and food industries, where large amounts of steam are required.

Process steam demand and electricity demand change independently according to season. When there is an imbalance between process steam demand and electricity demand, the back pressure turbine cannot respond to this imbalance by itself, and this imbalance is adjusted by power supply increase or decrease from the network or by reduced pressure and temperature steam from the HP steam source.

Saved heat (E) by the back pressure turbine is shown in the following formula if the total heat of the steam is used effectively.

(2.14) E = H t η i G ( H ′ t η i t ) h c

where

G: steam flow (kg/h),

η i : internal efficiency of back pressure turbine,

η it : internal efficiency of condensing turbine,

H t : adiabatic heat drop of back pressure turbine (kJ/kg),

H t ′ : adiabatic heat drop of condensing turbine (kJ/kg),

h c : potential heat of steam in condenser (kJ/kg),

Saved heat E is shown in Fig. 2.14 relating to turbine inlet pressure, temperature, and turbine outlet pressure. This figure indicates that saved heat will be raised by larger H t with higher inlet temperature and pressure and/or lower back pressure. The same logic can be applied to an extraction back pressure turbine.

Figure 2.14. Economic efficiency of the back pressure turbine.

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Condenser Flood and Vacuum Tightness Tests

Dipak K. Sarkar , in Thermal Power Plant, 2017

Abstract

A condenser creates the lowest possible turbine back pressure by evacuating air from a condenser shell. A liquid ring vacuum pump and/or steam-jet air ejector is used to maintain vacuum in a condenser. Heat liberated from turbine exhaust steam is removed by either water or air, or a combination of these. Water cooling is either the "recirculating" type or "once through" type. A condenser may be a "single-pressure" or "multipressure" type. Air ingress causes fall in condenser vacuum with consequent rise in turbine heat rate. Condenser flood and vacuum tightness tests identify locations through which air ingress takes place. Eliminating these defective locations ensures a vacuum-tight condenser. Heat Exchange Institute recommends evacuation of air from atmospheric pressure to 33.86   kPa absolute pressure Hg in about 1800   s.

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Boiler plant and auxiliaries

In The Efficient Use of Energy (Second Edition), 1982

Reducing Turbines

It is clear that when using a back-pressure turbine the steam available for process and the power supplied by the turbine are directly related. Most installations however require to be able to vary these quantities independently of each other, and to meet this requirement reducing turbines were introduced. These consist of a back-pressure turbine to which a low-pressure turbine is added, the back-pressure turbine being designed to pass the full quantity of steam required for process. As the demand for process steam is reduced, automatic governor gear bypasses the excess steam into the low-pressure turbine and at the same time the supply of high-pressure steam is reduced. As the turbine is also supplied with the usual governor gear to maintain speed approximately constant with varying load, the load and the process steam may be varied independently of one another, and the process steam supply maintained at almost constant pressure.

The economical results which can be obtained when using this plant for both electrical power and process steam have resulted in an increased demand for this type of turbine in recent years. In most cases the heat transferred to process is derived almost entirely from the latent heat of the steam. 28°C superheat at the turbine exhaust is usually desirable, so that the steam will reach the heater or plant in a dry condition but without excess temperature.

A general tendency in recent turbine development has been in the direction of increasing the initial temperature at which steam is supplied to the turbine; this results in increased thermal efficiency.

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Pass-out and Back-pressure Turbines—The Throttling Calorimeter

Wilfrid Francis , Martin C. Peters , in Fuels and Fuel Technology (Second Edition), 1980

Publisher Summary

This chapter discusses the pass-out and back-pressure turbines. The maximum thermal efficiency possible from a turbine operating on a modified Rankine cycle is about 40%. In practice, departures from the ideal cycle and other losses reduce this value, so that the efficiency of modern base load power stations is about 33%. The average for power stations in Great Britain is about 25%. By taking steam from the turbine before condensing, the whole of this steam is available for process work, or for heating, at the temperature and pressure at the exit to the turbine. As, under these circumstances, the bulk of the heat of the steam is usefully employed, the overall efficiency of the system is high and may reach 80% when all the exhaust steam can be usefully employed. If only a proportion of the steam passing through the first stages of the turbine is used for process heat, the unit is called a pass-out turbine. If all the steam is used for process heat, the unit is called a back-pressure turbine. The proportion of heat converted into electricity is calculated from the enthalpy drop in the turbine before pass-out is affected.

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Introduction

Swapan Basu , Ajay Kumar Debnath , in Power Plant Instrumentation and Control Handbook (Second Edition), 2019

2.7.2 Pass Out Turbine

There are some cases where the power available from a back-pressure turbine is less than required by the plant (Fig. 1.32). This may be due to a low heating requirement, a high back pressure, or a combination of both. The problem may be overcome in a two-stage turbine, where the main incoming steam expands through the high-pressure stage and supplies the heating steam from the exhaust. The balance steam passes through the low-pressure stage of the turbine to satisfy the power requirement. For any further heating steam requirement at any other pressure and temperature, a number of stages may be incorporated in the turbine for optimum power and heating output.

Fig. 1.32

Fig. 1.32. Schematic diagram of pass-out turbine.

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Turbine plant systems

British Electricity International, in Turbines, Generators and Associated Plant (Third Edition), 1991

1.7 Boiler feed pump turbine governors

On CEGB 660 MW turbine-generator systems, it is normal practice to use an auxiliary back-pressure turbine to drive the main feed pump. The feed pump is required to maintain the boiler drum level during steam raising prior to start-up of the main turbine, and also over the full load range of the main turbine. The precise arrangements for achieving feed flow control are outside the scope of this chapter, but the part played by the boiler feed pump turbine (BFPT) is to drive the pump at a speed such as to maintain a constant pressure drop across the feed regulating valves which follow the pump. The feed flow may then be controlled by setting the feed valve position. This requires a variable-speed turbine whose speed or steam valve position demand may be set by the feed control system.

The normal exhaust route for the BFPT is to the IP/LP crossover pipe of the main turbine, although at low loads it will be routed through to the condenser, automatic changeover from one exhaust to the other being sensed by a load-dependent parameter of the main turbine — usually crossover pressure.

The steam supply to the BFPT may be derived from one or other of two sources, or from a mixture of the two. Normally, when the main turbine is on-load, the second-stage blading of the BFPT is supplied with steam from the main turbine HP exhaust via bled-steam emergency and governing valves. Either one pair, or two pairs in parallel, of these valves are fitted. Because of non-availability of steam from the main turbine HP exhaust prior to start-up and when the main turbine is operating at low loads, the first-stage blading of the BFPT may be supplied with boiler steam through a single pair of live steam emergency and governing valves. In the event of non-availability of the bled-steam supply at full-load on the main turbine, the live steam valves will supply sufficient steam to enable full-load to be maintained.

The nominal maximum power rating of such a boiler feed pump turbine is about 15 MW and the possibility of a load rejection leading to a potential overspeed must be considered. The loss of load could occur as a result of loss of feed pump suction. If this occurred with the turbine speed and power output close to the full-load value, an extremely rapid acceleration would result due to the low rotor inertia. In the unlikely event that the turbine did overspeed to destruction, the strength of the casing would contain the disaster and protect personnel. However, it is normal practice to follow main turbine practice and fit the BFPT with a three-channel electronic governor, coupled with separate overspeed protection.

The turbine governor is a modular arrangement using many of the same elements applied to the main turbine. Thus, speed-sensing and on-load testing facilities will be similar in principle. The emergency valves upstream of the governing valves provide protection shutdowns of the BFPT in the event of failure of more than one governor channel or in the event of a mechanical trip condition (e.g., loss of lubricating oil) being sensed. The valve relay systems use similar servo-valves to those fitted to the main turbine and the following stages of hydraulic operation are often similar in principle to those of the main turbine. It has sometimes been possible to omit the pilot stage associated with the main turbine valve relays and to drive the power piston direct with the servo-valve. Referring back to Fig 2.24, this would be equivalent to considering the relay to comprise only the servo-valve and primary ram. In all cases, the hydraulic fluid is supplied from the main turbine system.

Live and bled-steam governing valves are controlled in parallel, with an offset being applied to the live steam valves so that they only begin to open when the bled-steam valves are fully open on a speed loop droop of about 5%. Any additional speed error, requiring further steam supply, progressively opens the live steam valves until they are also fully open.

During start-up of the main turbine, the steam is supplied via the live steam valve. A common means of ensuring this automatically is to superimpose on the bled-steam channel a limited maximum opening characteristic as a function of bled-steam pressure, as shown in Fig 2.32. This will keep the bled-steam valves closed until the HP exhaust pressure has risen to the value at which it can sustain the feed flow demand via the BFPT.

FIG. 2.32. Valve and pressure sequencing diagrams for the BFPT

These features and others, engineered by techniques similar to those of the main turbine governor, enable convenient interfaces to be provided for automatic feed control systems and automatic run-up.

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District Heating with Combined Heat and Electric Power Generation

Richard H. Tourin , in Advances in Energy Systems and Technology, Volume 1, 1978

E Hot Water versus Steam District Heating Systems

At this point it is important to discuss the advantage of hot water district heating used in Europe over the present United States steam systems. There are two major advantages to a hot water district heating system: improved system control (including load leveling) and increased fuel efficiencies (for combined heat and power generation).

Control of the quantity of heat which reaches the consumer in a hot water system is achieved by the control of both flow rate and temperature. These two parameters are routinely monitored and controlled at the central heat plant in response to the consumer's heat requirement (based on ambient temperature) and electrical loading. The hot water system allows for greater flexibility in matching electrical load to generating capacity. The heat energy can be stored in hot water tanks during periods of high electrical demand, and sent out during periods of low demand. A system with these capabilities can level the system heat demand, while continuously following the electrical load (Muir, 1973,1975).

To illustrate the fuel savings obtainable by using a hot water distribution system rather than a steam system, calculations have been made for the systems shown in Figs. 10 and 11. The hot water combined heat and power generation system is represented schematically in Fig. 10. Figure 11 represents our model of a steam combined heat and power generation system. Calculated data for the two models are given in Table II. Each system supplies 1,000,000 Btu of useful heat energy and 110.02 KW h of electric energy. Column 1 in Table II illustrates that 24% more fuel is required by the steam system to supply the same power requirements as the hot water system.

Fig. 10. Water system schematic for comparison of water and steam systems (see Fig. 11).

Fig. 11. Steam system schematic for comparison of water and steam systems, (a) Back-pressure steam system, (b) Feedwater heating for (a). (c) Conventional unit to supplement (a) and (b).

TABLE II. Comparison of Water and Steam Systems

Location 1 2 3 4 5 6 7 8 9 10 11
I. Water system
  a. Substance a F – S S W H H – – H E
  b. Quantity (lb) – – 1080 1080 1080 – – – – – –
  c. Pressure psi (absolute) – – 1500 17.2 17.2 – – – – – –
  d. Temperature (°F) – – 1000 220 220 – – – – – –
  e. Enthalpy (Btu/lb) – – 1490 1142 188 – – – – – –
  f. Heat content (K Btu) 1654 – 1609 1234 203 1031 31 1000 – 1000 110.02
  g. Electricity (kW h) – – – – – – – – – – –
IIA. Steam system
  a. Substance a F – S S W – S S W H E
  b. Quantity (lb) – – 905 905 905 – 90 815 815 – –
  c. Pressure [psia (absolute)] – – 1500 200 – – 200 200 – – –
  d. Temperature (°F) – – 1000 549 220 – 549 549 100 – –
  e. Enthalpy (Btu/lb) – – 1490 1295 188 – 1295 1295 68 – –
  f. Heat content (K Btu) 1387 – 1349 1172 170 – 117 1055 55 1000 51.66
  g. Electricity (kW h) – – – – – – – – – – –
IIB. Feed heating for IIA
  a. Substancea a F W S S W W – – – – E
  b. Quantity (lb) – 905 161 161 161 905 – – – – –
  c. Pressure [psi (absolute)] – – 1500 17.2 – – – – – – –
  d. Temperature (°F) – 50 1000 220 220 220 – – – – –
  e. Enthalpy (Btu/lb) – 18 1490 1142 188 188 – – – – –
  f. Heat content (K Btu) 247 16 240 184 30 170 – – – – –
  g. Electricity (kW h) – – – – – – – – – – 16.40
IIC. Conventional unit to supplement IIA and IIB
  a. Substance a F – – – – – – – – – –
  f. Heat content (K Btu) 420 – – – – – – – – – –
  g. Electricity – – – – – – – – – – 41.96
Total for IIA, IIB, and IIC 2054 – – – – – – – – 1000 110.02
a
F = Fuel, W = water, S = steam, E = electricity, H = heat.

The hot water system model used here is a straightforward application of combined heat and power technology to a back-pressure turbine scheme (see columns 1–11 which refer to Fig. 10 under Water system, Table II). The steam system model, however, is complicated by the need for feedwater heating (no condensate is returned to the steam combined heat and power generation plant), and by the need for additional electric generation to match the output of the hot water system. (This power output difference is made up by a conventional condensing-type electric generating plant.) Thus, Fig. 11 consists of the following:

(a)

A straightforward back-pressure steam system which supplies 1,000,000 Btu of useful heat to the distribution system while generating electric power (but substantially less power than the hot water system supplying the same quantity of heat) (IIA in Table II).

(b)

A back-pressure steam system which supplies heat for boiler feedwater while producing a small quantity of electricity. (At the steam combined heat and power generation plant, this steam and electric generation would be integrated with the system of (a). It is separated here to demonstrate the fuel requirements for feedwater heating which are associated with the steam system, which does not incorporate condensate return) (IIB in Table II).

(c)

A conventional condensing power plant which supplies the electric power required to match steam system output to the higher electric output of the hot water system (IIC in Table II).

The data associated with each component of this steam system model are displayed in Table II under the appropriate column numbers (which refer to the schematics of Fig. 11). To summarize the numbering scheme of Table II, we have:

1.

Fuel input to boiler.

2.

Water input to boiler feedwater makeup

3.

Steam from boiler output.

4.

Steam exhausted from back pressure turbine.

5.

Boiler feedwater.

6.

Boiler feedwater makeup.

7.

Heat lost in district heating distribution system.

8.

Heat input to customer's heating system.

9.

Condensate discharged to sewer.

10.

Heat used in customer's heat system.

11.

Electric energy generated.

In the calculations for Table II a boiler efficiency of 85% has been used. Turbine efficiency to electric power has been taken as 80% of the maximum theoretically obtainable by expansion to the pressure indicated, with the balance of the heat appearing in the exhaust steam. For simplicity in calculation only single-stage feedwater heating has been assumed, hot water heating in system I has been assumed to be single stage, and auxiliary power requirements have been ignored.

It may be noted that in the above example the water and steam systems are brought to equal outputs, for comparative purposes, by adding the production of electric power in a conventional system (IIC). This implies that the heat load is limited in size and that the electric load will never be limited. This will always be the case when there is a connection to the grid of adequate capacity. For isolated systems (no grid connection) the higher fuel used by the steam system is only about 10%.

The major reasons for the higher fuel consumption for the steam system are (1) the need to generate more electric power in conventional units, (2) higher distribution losses (10% versus 3%), and (3) loss of heat by the customer in the condensate.

There are other less dramatic advantages to the hot water system including the following:

1.

Hot water has economically been distributed, at a constant pressure, as far as 60 km (37 miles) with pumping power requirements of only 0.5% to 3% of the system's thermal power capacity. This permits greater flexibility in scheduling heat delivery from the most economical stations at times of low load. By contrast, steam distribution is practicable only up to distances a mile or two from the steam plant.

2.

The simplicity of a low-pressure integrated hot water system affords great system reliability.

3.

Steam metering is much more difficult than hot water metering, causing greater amounts of unaccounted-for steam.

These are the factors that have caused the longevity and continuing growth rates for the European hot water systems. The greater fuel efficiencies of hot water systems can reverse the unfavorable economics for district heating in the United States, considering the escalating cost of fuel. The interested reader will find a more detailed comparison of steam and hot water district heating systems in Muir (1975).

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